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STEAM POWER PLANT 
AUXILIARIES AND ACCESSORIES 

TERRELL CROFT Editor 



CONTRIBUTORS 

The following staff engineers of the Terrell Croft Engineering Company have 
contributed manuscript or data or have otherwise assisted in the preparation 
of this work: 

Edmond Siroky, Head Mechanical Engineer 
A. J. Dixon E. R. Powell 

I. V. LeBow I. O. Royse 

Also the following engineers furnished manuscript or data for the respective 
divisions the titles of which follow their names: — 

A. C. Staley — Condensers and Methods of Recooling Condensing Water 
Julius Wolf — Injectors 



BOOKS BY 

TERRELL CROFT 

PUBLISHED BY 

McGRAW-HILL BOOK COMPANY, Inc. 

The American Electricians' Handbook, 
Flexible Leather, 7 X 41, 712 Pages, 
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Flexible Cover, Pocket Size, 448 Pages, 
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Cloth, 8 X 5^ 318 Pages, 302 Illus- 
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Practical Electricity, 

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Central Stations, 

Cloth, 8 X 5§, 330 Pages, 306 Illus- 
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Cloth, 8 X 5§, 412 Pages, 514 Illus- 
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STEAM POWER PLANT 
AUXILIARIES AND ACCESSORIES 



TERRELL CROFT, Editor 

CONSULTING ENGINEER. DIRECTING ENGINEER, TERRELL CROFT ENGINEERING CO. 

MEMBER AMERICAN INSTITUTE OF ELECTRICAL ENGINEERS. 

MEMBER AMERICAN SOCIETY OF MECHANICAL ENGINEERS. 

MEMBER OF THE ILLUMINATING ENGINEERING SOCIETY. 

MEMBER AMERICAN SOCIETY FOR TESTING MATERIALS. 



First Edition 
First Impression 



McGRAW-HILL BOOK COMPANY, Inc. 
NEW YORK: 370 SEVENTH AVENUE 

LONDON: 6 & 8 BOUVERIE ST., E. C. 4 

1922 









C>1 



c 



*" 



Copyright, 1922, by Terrell Croft 



M-l 



THE MAPLE PRESS YORK PA 

MAY 1 1 1922 

©CU674238 






PREFACE 

Most of the preventable losses in the engine rooms of steam 
power plants occur in connection with the auxiliary equipment. 
Generally speaking, there is not a great deal that the operating 
engineer can do to increase the efficiencies of the prime movers 
— the turbines or engines. It is also a fact that, as a rule, 
the prime movers in a plant give relatively little trouble and 
involve relatively little maintenance expense. Most of the 
trouble and maintenance expense is due to the auxiliaries. 
Thus, it follows that, in a sense, the auxiliary equipment com- 
prises the most important part of that portion of the power- 
plant equipment which transforms the heat in the steam into 
power. 

Hence, in this book, it has been the endeavor to give such 
data as will enable the operator to select, and properly install, 
auxiliary equipment which will insure the generation of power 
at the least cost. Furthermore — and quite as important — it 
has been the aim to provide the information whereby this auxil- 
iary equipment can be so operated and maintained that its 
preventable losses will be a minimum and that its up-keep 
expense will be as small as is feasible. 

Drawings for all of the 411 illustrations were made especially 
for this work. It has been the endeavor so to design and 
render these pictures that they will convey the desired infor- 
mation with a minimum of supplementary discussion. 

Throughout the text, principles which are presented are 
explained with descriptive expositions or with worked-out 
arithmetical examples. At the end of each of the 13 divisions 
there are questions to be answered and, where justified, 
problems to be solved by the reader. These questions and 
problems are based on the text matter in the division just 
preceding. If the reader can answer the questions and 
solve the problems, he then must be conversant with the sub- 
ject matter of the division. Detail solutions to all of the 
problems are printed in the appendix in the back of the book. 



viii PREFACE 

As to the method of treatment : Pumps are first considered 
because almost every power plant, regardless of size, requires 
pumps of some sort, for its operation. Hence, there are 
divisions on pump calculations, direct-acting steam pumps, 
crank-action pumps, centrifugal and rotary pumps. Next 
follows a discussion of boiler-feeding apparatus such as boiler- 
feed pumps and their governors, injectors, and gravity boiler- 
feeding devices. The problems of feed-water heating are then 
treated in the divisions on feed-water heaters and economizers. 

Following this are divisions on condensers and methods of 
recooling condensing water which, it is believed, are, both 
economically and practically, very thoroughly treated. 
Finally, the divisions on steam piping, live- and exhaust- 
steam separators, and steam traps explain how these elements 
should be selected, installed, and maintained. They also 
present solutions to the problems of preventing losses from 
and in steam pipes. 

With this, as with other books which have been prepared by 
the author, it is the sincere desire to render it of maximum 
usefulness to the reader. It is the intention to improve the 
book each time it is revised and to enlarge it as conditions may 
demand. If these things are to be accomplished most effec- 
tively, it is essential that the readers cooperate with us. This 
they may do by advising the author of alterations which they 
feel it would be advisable to make. Future revisions and 
additions will, insofar as is feasible, be based on such sugges- 
tions and criticisms from the readers. 

Although the proofs have been read and checked very care- 
fully by a number of persons, it is possible that some undis- 
covered errors may remain. Readers will confer a decided 
favor in advising the author of any such. 

Terrell Croft. 
University City, 

St. Louis Mo., 



ACKNOWLEDGMENTS 

The author desires to acknowledge the assistance which 
has been rendered by a number of concerns and individuals 
in the preparation of this book. 

Considerable of the text material appeared originally as 
articles in certain trade and technical periodicals among 
which are: Power, National Engineer, Power Plant Engineer- 
ing, and Southern Engineer. 

The author is particularly indebted to Mr. H. H. Kelley and 
to Mr. F. A. Burg, manager of the condenser section of the 
Westinghouse Electric and Manufacturing Company for their 
contributions to the condenser division. Acknowledgment is 
also here given to Mr. F. F. Nickel for his able assistance in 
the matter on pumps. 

Among the manufacturers who cooperated in supplying 
text data and illustrations are: The Cooling Tower Company; 
Worthington Pump and Machinery Corporation; Union Steam 
Pump Company; The Goulds Manufacturing Company; Schutte 
and Kcerting Company; H. S. B. W .-Cochrane Corporation; 
Green Fuel Economizer Company; B. F. Sturtevant Company; 
Westinghouse Electric and Manufacturing Company; C. H. 
Wheeler Manufacturing Company; Wheeler Condenser and 
Engineering Company; Spray Engineering Company; Crane 
Company. 

Special acknowledgment is hereby accorded Edmond Siroky, 
Head Mechanical Engineer of The Terrell Croft Engineering 
Companjr, who has been responsible for the technical accuracy 
of the book. 

Other acknowledgments have been made throughout the 
book. If any has been omitted, it has been through oversight 
and, if brought to the author's attention, it will be incorporated 
in the next edition. 

Terrell Croft. 



CONTENTS 



STEAM POWER PLANT AUXILIARIES AND ACCESSORIES 

BY 

Terrell Croft 

Page 

Frontispiece iv 

Preface vii 

Acknowledgments ix 

List of Symbols xii 

Division 1. — Pump Calculations 1 

Division 2. — Direct-Acting Steam Pumps 39 

DrvisioN 3. — Crank-Action Pumps 75 

Division 4. — Centrifugal and Rotary Pumps 101 

DrvisioN 5. — Injectors 155 

DrvisioN 6. — Boiler-Feeding Apparatus (Pump Governors) . . 171 

DrvisioN 7. — Feed-Water Heaters 207 

Division 8. — Fuel Economizers. . 251 

Division 9. — Condensers 277 

DrvisioN 10. — Methods of Recooling Condensing Water. . . 329 

DrvisioN 11. — Steam Piping of Power Plants 363 

Division 12. — Live-Steam and Exhaust-Steam Separators . . . 385 

Division 13. — Steam Traps 403 

Solutions to Problems 415 

Index 425 



XI 



STEAM POWER PLANT AUXILIARIES AND ACCESSORIES 

List Of Symbols 

The following list comprises practically all of the symbols which are used 
in formulas in this book. Symbols which are not given in this list are 
defined in the text where they are first used. When a symbol is used with 
a meaning different from that below, the correct meaning is stated in the 
text where the symbol occurs. 

Section 
Symbol Meaning First Used 

A Piston area, in square inches 21 

Abh Area of boiler heating surface, in square feet 192 

A i Internal area of pipe, in square inches 440 

A f Area, in square feet 19 

C g Specific heat of combustion-gases 303 

C w Specific heat of water 303 

d Diameter of impeller, in inches 121 

di Internal diameter of pipe, in inches 19 

di m Inside diameter of main pipe, in inches 444 

d External pipe-diameter, in inches 448 

d p Piston-diameter, in inches 26 

d s Steam-piston-diameter, in inches 28 

D Density of steam, in pounds per cubic foot 440 

Di Density, in pounds per cubic inch 21 

D c Duty, in foot pounds per 100 pounds of coal 47 

Dfc Duty, in foot pounds per 1,000,000 B.t.u 49 

T> s Duty, in foot pounds per 1,000 pounds of steam 48 

ei Coefficient of linear expansion 447 

E Efficiency in per cent , 392 

E>i Hydraulic efficiency, in per cent 37 

Ei Indicated efficiency, in per cent 35 

E m Mechanical efficiency, in per cent 41 

E m Efficiency of motor, in per cent 138 

E p Efficiency of pump, in per cent 138 

E t Total efficiency, in per cent 42 

E t Thermal efficiency, expressed decimally 321 

E v Volumetric efficiency, in per cent 25 

E V d "Volumetric efficiency, expressed decimally 26 

g Acceleration due to gravity in feet per second, per second = 

32.2 7 

H Heat, in B.t.u 49 

H Total heat of steam, in British thermal units per pound 244 

H/ Per cent, saving in heat-content of fuel 244 

xii 



LIST OF SYMBOLS xill 

Section 
Symbol Meaning First Used 

H t Heat, in B.t.u. given up by the steam per hour 348 

H v Latent heat of vaporization of steam 189 

/ Current, in amperes 138 

K Condensation, pounds per hour per square foot of pipe surface . 499 

K A constant 107 

K n A constant 406 

I Linear expansion of pipe, in inches 447 

L Length of stroke, in inches 21 

L b Minimum pipe length required for bend 448 

L e Pipe-length, in inches, having resistance equivalent to one 

90-deg. elbow 446 

Lf Length, in feet 39 

L h Height, infeet 268 

L h Static head, in feet 5 

Lh/c Friction head, in feet, due to pump passages and valves 9 

L h ff Friction head, in feet, due to pipe bends 9 

L h /i Fricton head, in feet, due to inlet flow 9 

Lhf P Friction head, in feet, due to straight pipe 9 

L h fT Total friction head, in feet 9 

Lhfv Friction head, in feet, due to valves in piping 9 

Lhmd Measured head, in feet, due to delivery lift 11 

Lh ms Measured head, in feet, due to suction lift 11 

Lhmp Head, in feet, due to back pressure on delivery-pipe outlet. . . 11 

Lhmt Total measured head, in feet 11 

L h T Total head, in feet 12 

L hu Useful head, in feet 34 

L hv Velocity head, in feet 7 

L p Length of pipe-line, in feet 448 

Lt Piston-travel, in feet per minute 26 

L v Pipe-length, in inches having resistance equivalent to one 

globe valve 445 

L w Width of belt, in inches 145 

M Relative humidity of the air expressed decimally 398 

N Revolutions per minute 118 

N s Number of strokes per minute 21 

P Pressure, in pounds per square inch 5 

P a Absolute pressure, in pounds per square inch 189 

Pbhp Driving horse power 41 

Pshp Boiler horse power 229 

Pd Discharge pressure, in pounds per square inch 49 

P D Hydrostatic pressure head, in pounds per square inch 49 

x 3.1416 19 

Phmv Vacuum, in inches of mercury 322 

Phmb Barometer reading, in inches of mercury 327 

Pi Intake pressure, in pounds per square inch 49 



XIV 



LIST OF SYMBOLS 



Symbol 

Pm 
Ps 

"uhp 
Pv 
P W 

*whp 

T 

T f 
Ty 

T f a 



Me A] 



Section 
First Used 



T f a 

Tf C 

Tfa 
Tfa 

T fg 

T fi 
Tfs 
T fs 
Tf w 
T f W 
T} w 
T'f W 

U 

v 

V 

V 

Va 
Vcf 

v gm 

I 

w 



Mean effective pressure, in pdffhds per square inch 322 

Steam pressure, in pounds per square inch. .' . 28 

Useful hydraulic horse power 34 

Vapor pressure, in inches- of mercury 398 

Total head-pressure, in pounds per square inch 28 

Actual hydraulic horse power. 35 

Absolute temperature, on Fahrenheit scale ,.' 321 

Temperature, in degrees Fahrenheit. 244 

Temperature change, in degrees Fahrenheit 447 

Average temperature, in degrees Fahrenheit, of water leaving 

cooling-tower f 419 

Temperature of air, in degrees Fahrenheit f . . 452 

Temperature of condensate, in degrees Fahrenheit 344 

Final temperature of condensed steam, in degrees Fahrenheit 189 
Dry-bulb-thermometer temperature, in degrees Fahrenheit . . 406 

Loss of gas temperature, in degrees Fahrenheit. . 303 

Temperature of intake water to injector, in degrees Fahrenheit 189 

Temperature of steam used for heating feed-water 266 

Temperature of steam, in degrees Fahrenheit j£ . . . 189 

Wet-bulb-thermometer temperature, in degrees Fahrenheit . . 392 

Temperature gain, water, in degrees Fahrenheit 303 

Temperature, of feed-water in degrees Fahrenheit 309 

Temperature of feed water, in degrees Fahrenheit, at exit of 

economizer ' 309 

Coefficient of heat transfer in B.t.u. per hour, per degree 

, Fahrenheit temperature difference , 277 

Velocity, in feet per second 7 

Volts /. ....... 138 

Volume of condenser, in cubic feet 342 

Volume, in cubic feet per minute 19 

Displacement, in cubic feet per minute 21 

Quantity of water, in gallons per minute 19 

Velocity, in feet per minute 19 

Weight of liquid, fa pounds 31 

Weight, in poinds per minute 21 

Weight of condensation, in pounds per hour 452 

Weight of coal, in pounds '. 47 

Weight of feed-water' entering heater, in pounds per hour . . . 262 

Weight of feed water leaving heater in pounds per hour 266 

Weight of gas, in pounds, per pound of coal burned 303 

Weight of moisture in steam, in pounds 476 

Steam rate, in pounds per hour 262 

Weight of steam, in pounds 48 

Pounds of water pumped per pound of steam 189 

Useful work, in foot pounds 31 



LIST OF SYMBOLS xv 

-j Section 

Symbol Mea^|ng. First Used 

W w •' Weight of water, in pounds. . . ?• 189 

W w Weight of water evaporated, per pound of -coal burned, in 

pounds 303 

W w Water rate, in pounds per hour 344 

W» -.Weight of water evaporated, in pounds per square foot per hour 398 

W^.^r Weight of water per boiler horse power per hour 229 

X Slip, in per cent . ! 24 

X Saving, in per cent •. + 309 

X Quality of steam 189 

X . Ratio 303 



x 






V 



STEAM POWER PLANT 
AUXILIARIES AND ACCESSORIES 



■Discharge 
Pipe 



DIVISION 1 

PUMP CALCULATIONS 

1. The Height To Which Water May Be Drawn By Pump- 
Suction depends principally: (1) Upon the condition of the 
pump as regards the 
tightness of its valves, 
piston- or plunger- 
packing and piston- 
rod packing . (2) 
Upon the water-fric- 
tion in the suction 
pipe (Sec. 8) and fit- 
tings. (3) Upon the 
temperature of the 
water. (4) Upon the 
altitude above sea-level. 



Note. — The practical 
maximum suction-lift is 
about 22 feet. 

EXPLANATI ON. A t - 

mospheric pressure at 
sea-level is about 14.7 
lb. per sq. in., absolute. 
A 2.31-ft. height of water- 
column is the equivalent 
of 1 lb. per sq. in. pres- 
sure. On this basis the 
theoretical suction-lift at 
sea-level is 14.7 X 2.31 
= 34 ft., nearly. But 
in actual practice a lift 




Foof'-V6i!ve--— 



Fig. 1. 



-How A Double-Acting Suction Pump Operates. 

of 22 ft. under sea-level atmospheric pressure is, due to unavoidable 
leakage, friction and vaporization (Sees. 5, 8, and 10), seldom exceeded. 

1 



STEAM POWER PLANT AUXILIARIES 



[Div. 1 



Therefore (Table 2), a pump lifting 22 ft. at sea-level would, at 1 mile 
above sea-level, where the atmospheric pressure is about 12.02 lb. per 
sq. in., give a lift of 12.02 X 22 -^ 14.7 = 17.9 ft. 

Note. — The net suction-lift of a reciprocating pump is the vertical 
distance, Lh S (Fig. 1), from the level of the water in the well, or other 
source of suction-supply, to the level of the discharge-valve seats. The 
total suction-lift comprises the net lift and the friction head (Sec. 6) due 
to water-friction. 

2. Table Showing Practical Pump Suction Lifts At Various 
Altitudes. — Ordinary atmospheric temperature is assumed. 
(Goulds Catalogue.) 



Altitude above sea-level 


Barometric pressure 


Practical 
suction 


Miles 


Feet 


Pounds 
per sq. in. 


Head in ft. 
of water 


lift of 

pumps, 

feet 


Sea-level 
H 

y 2 

H 
l 

IK 
IH 

2 


Sea-level 
1320 
2640 
3960 
5280 
6600 
7920 
10560 


14.70 
14.02 
13.33 
12.66 
12.02 
11.42 
10.88 
9.88 


33.95 
32.38 
30.79 

29.24 
27.76 
26.38 
25.13 

22.82 


22 
21 
20 
18 
17 
16 
15 
14 



In The Pumping Of Hot Water the tendency of the water 

to vaporize under different de- 
grees of absolute pressure must 
be considered. As the suction- 
lift of a pump increases (Fig. 
2), the maximum temperature 
of the water that can be pumped 
decreases. Generally, it will be 
found practically impossible to 
lift water at a temperature 
above 150 deg. fahr. Hence, 
where a boiler feed-pump (Fig. 
3), receives its suction supply 

from an open feed-water heater, the water must flow to the 

pump under (Sec. 4) a static head. 




IZW864Z0Z468 lZ'.Jt ZO 24 28 32 3^ 
Intake Head in) Suction Lift in Feet 
Feet 

Fig. 2. — Diagram Showing Practi- 
cal Intake Pressures At Different 
Temperatures; Also The Theoretical 
Water Lift Of A Pump. Calculations 
Are For Sea Level. 



Sec. 4] 



PUMP CALCULATIONS 



4. The Static Head Of A Fluid Column, as a column of water 
(L h2 , Fig. 4) in a standpipe, is the vertical distance between 
the base and the top surface of the column. It is understood 
to mean the pressure which the column imposes on the plane 
which is taken as a base. Thus, a 30-ft. static head means the 



Cold Wetter Inlet- --> 



Pump Exhaust- Pipe 
Connection- 



"m^^^mm^m^^^^^^^. 







Fig. 3.— Boiler Feed-Pump Taking Water-Supply From Open Feed-Water Heater. 

pressure, per unit of base area, which is due to the weight of a 
fluid column 30 ft, high. 

Explanation.— L hl (Fig. 4) is the static head of the column of water 
above the plane AB, while L h2 is the static head above the plane IF. 
A column of water 1 in. square and 1 ft. high weighs, approximately, 
0.433 lb. Hence, the static heads, L hx and L h2 , may be readily trans- 
lated into terms of pressure. Thus, if L hl = 40 ft., then the pressure on 



STEAM POWER PLANT AUXILIARIES 



[Div. 1 



AB = 0.433 X 40 = 17.32 lb. per sq. in. If L h2 = 50 ft., then the 
pressure on XY will be: 0.433 X 50 = 21.65 lb. per sq. in. 

Note. — The Inlet Static Heads (Inlet Pressures) for Boiler 
Feed-Pumps (L h , Fig. 3) drawing water from open feed- water heaters 
should, in order to secure satisfactory service, be from about 1.5 ft. for 
water at 165 deg. fahr. to about 11.5 ft. for 
water (Fig. 2) at 210 deg. fahr. The pressures 
due to these heads are necessary to counteract 
the tendency of pumps to become steam-bound, 
or filled with vapor from heated water. When 
a reciprocating pump is in this condition, the 
piston traverses the cylinder without producing 
a discharge. The vapor in each end of the cyl- 
inder is compressed during one stroke and re- 
expands during the opposite stroke, while the 
boiler-pressure above the delivery valves holds 
them seated. 




Lh, L h . 



Fig. 4. — Illustrating Static 
Head Of A Liquid. 



5. The Head Or Pressure Due To A 
Column Of Water May Be Converted 
Into Equivalent Terms Of Unit Pressure by the following 
formula : 

L h 



(1) P = 0.433L A = 2 



(pounds per square inch) 



Wherein P = pressure, in pounds per square inch. Lh = 
static head, in feet. 

Example. — A direct-acting steam pump (Fig. 1) is discharging into 
an open tank. What is the pressure, due to the discharge head, on the 
discharge-end of the pump-plunger if the vertical distance from the 
horizontal axis at the pump-cylinder to the level of the water in the tank 
is 25 feet? 

Solution.— By For. (1) P = L h /2.S1 = 25 -r- 2.31 = 10.8 lb. per 
sq. in. 

6. A Pump Must Overcome Certain Resistances And 
Pressures in delivering water or other liquids. The following 
must be considered in calculations: 

(1) Velocity head or velocity pressure, which is the head or 
pressure required to set the liquid in motion and give it the 
velocity which it will have at the final stage of its movement. 

(2) Friction head or friction pressure, which is the resistance 
head or pressure required to overcome the resistance due to 



Sec. 7] PUMP CALCULATIONS 5 

the friction between the liquid and the surfaces of the pipes, 
fittings, valves and pump-passages through which it flows. 

(3) Measured head or measured pressure, which is the vertical 
height, or the equivalent pressure due to this height, from a 
lower to a higher plane in the pumping system. The lower 
plane may be the surface of a cooling pond. The higher plane 
may be the center of the mouth of the discharge pipe which 
conveys the water into a tank. 

Note. — The Dynamic Head Or Pressure is the sum of the velocity- 
head and friction-head. 

7. The Velocity Of A Liquid In A Pipe Must Be Produced 
By Pressure. The pressure may be thought of as the pressure 
which is produced (Sec. 5) by a vertical column of the liquid. 
If friction and all other resistances are neglected, the velocity 
produced by a certain head will be equivalent to the velocity 
attained by a falling body which descends a distance equal 
to the head. See also Div. 4. It can be shown that: 

(2) v = -\/2g Lhv (feet per second) 
Wherein v = velocity, in feet per second, g = acceleration 
due to gravity, in feet per second per second = 32.2 approxi- 
mately. Lhv = head necessary to produce the velocity, in 
feet. 

If the velocity is known, the head to which it is due may 
be found by the above formula rearranged : 

(3) L hv = ^- (feet) 

Note. — As the velocity is often small, the hydraulic head necessary 
to produce it will be small. It is, therefore, often neglected. See 
following sections and examples. 

Example. — What velocity will result from a head of 50 ft. of water 
when all the head is available for imparting velocity to the water? 
Solution.— By For. (2) v = \/2gL hv = V2 X 32.2 X 50 = 56.7 
ft. per sec. 

Example. — What velocity head must a pump produce if it is to dis- 
charge a liquid at a velocity of 10 ft. per sec? Solution. — By For. (3) : 
L hv = v*/2g = (10) 2 v(2X 32.2) = 1.58 ft. 

8. The Friction-Head On A Pump may be necessary for 
overcoming the following resistances: (1) The friction (Tables 
14 and 15) due to the flow of a liquid through straight piper. 



6 



STEAM POWER PLANT AUXILIARIES 



[Div. 1 



(2) The friction of the liquid entering (Figs. 5, 6, and 7) the 
suction or inlet pipe. (3) The friction due to the flow through 
the pump-valves and passages within the pump. (4) The fric- 



Suction 
Pipe— 







Square Orifice - : 

Fig. 5. — Pump Suction. 
Pipe With Square En- 
trance Orifice. 




Coupling* 


■V 


I-''' 


■Suction 
Pipe 


s 


' 4 


!===* 


im 



Strainer- 



Fig. 6. — Pump Suction- 
Pipe With Strainer. 



■Furineh 



Fig. 7.— Funneled End Of 
Pump Suction-Pipe. 



tion due to the flow (Figs. 8, 9, and 10) through pipe-fittings; 
this resistance is caused by the change of direction of the 
flow, and by the roughness of the fittings. (5) The friction due 
to flow through valves in the piping; with gate valves this resist- 
ance is negligible. 



Direction of Flow-. 




Direction of Fhw-. 



'•Couplings 



Fig. 8. — Turn In Pump 
Piping Made With Long- 
Radius Bend. 




*,pe 


>m 


/ Tee--* 


/ 



'-•Direction 
of Flow— 



Fig. 9. — Turn In Pump- 
Piping Made With Elbow 
Having A Radius Equal 
To Pipe-Diameter. 



Fig. 10. — Sharp Turn 
In Pump Piping Made 
With Plugged Tee. 



Note. — The head due to friction of the water entering the suction 
pipe is called the entrance-head. 

9. The Total Friction-Head On A Pump is the sum of the 

resistances enumerated in Sec. 8. It may be expressed by the 
following formula: 



Sec. 10] 



PUMP CALCULATIONS 



(4) L hfT = L hfp + L hfi + L hff + L A/ „ + L hfc (feet) 

Wherein L A/r = the total friction head, in feet. L hfp = 
the friction-head, in feet, due to flow through straight runs 
of suction and discharge piping. L hfi = the friction-head, 
in feet, due to the inlet flow. L hff = the friction-head, in 
feet, due to pipe-fittings which change the direction of the 
flow. L hfv = the friction-head, in feet, due to valves in the 
piping. L hfc = friction-head, in feet, due to flow through 
passages and valves in pumps. 

10. The Measured Heads In Pump Operation, L Hh in 
Fig. 11 (see Sec. 6 for definition of Measured Head) com- 



discharge Level 




■ •; Intake- --■&> 



Fig. 11. — Illustrating Useful Pump- Work. 



prise the following: (1) The suction-lift of the water. The 
height of this lift is (L hms , Fig. 11) from the level of the dis- 
charge valve seat to the* surface of the suction-water. (2) 
The delivery-lift of the water. This (L hmdf Fig. 11) is mea- 
sured from the level of the seat of the discharge valve to the 
center of the outlet orifice of the delivery pipe, where the water 
issues horizontally from the pipe and falls by gravity. Or 



8 



STEAM POWER PLANT AUXILIARIES 



[Div. 1 



(Fig. 11) it is measured from the level of the discharge valve 
seat to the level of the water above the discharge orifice, 
where the outlet end of the pipe is submerged. (3) The 
head due to pressure, above atmospheric pressure, on the liquid 
in the vessel into which the delivery-pipe discharges. If water 
is being delivered to a boiler, this head is equivalent to the 
steam-gage pressure in the boiler. 

Note. — The Suction Lift Of A Centrifugal Pump is measured 
from the level of the water in the well to the center of the impeller. 




Fig. 12. — An Imperfectly Laid Suction-Line. 

Note. — When suction pipes are laid underground, in trenches, care 
should be exercised to run them in a slightly declining straight line 
toward the source of supply. High places or hummocks (Fig. 12) in 





W^^^^W^B- 



V- 'Rrvery. 
^Intake 

STW k cr/d "v \ 



".-Gate Valve 



?^>" ^^Intake-Pipe Inclining Toward Well-' " Intake'-X 




Fig. 13. — Suction Well Supplied Through Intake Pipe. 



the suction line afford pockets for the accumulation of air. Such air- 
pockets reduce the effective area of the pipe and cut down the water 
supply to the pump. 



Sec. 11] PUMP CALCULATIONS 9 

Note. — When the distance from a pump to a natural source of suction- 
supply, as a pond, lake, or stream, exceeds about 100 ft., it is advisable 
(Fig. 13) to sink a suction- well close to the pump. The intake-pipe 
should then incline toward the well. 



11. The Total Measured Head On A Pump is the sum of 

the measured heads enumerated in Sec. 10. It may be ex- 
pressed by the following formula: 

(5) LhmT — Lhms + Lhmd + Lhmp (feet) 

Wherein LhmT = the total measured head, in feet. Lhms = 
the measured head, in feet, due to suction-lift. Lhmd = the 
measured head, in feet, due to delivery-lift. L hmp = the 
measured head, in feet, due to steam, compressed air, or other 
fluid pressure in the vessel into which the delivery-pipe dis- 
charges. If the delivery pipe discharges freely into the 
atmosphere, then Lh mp is zero. 

12. The Total Head On A Pump is the sum of: the velocity- 
head (Sec. 7), the total friction-head (Sec. 9) and the total 
measured-head (Sec. 11). It may be expressed by the following 
formula : 

(6) LhT = L hv + L h fT + LhmT (feet) 

Wherein Lh T = the total head, in feet. Lhv = the velocity- 
head, in feet. LhfT = the total friction-head, in feet. LhmT = 
the total measured-head, in feet. 

13. The Friction Of Water In Straight Pipes Is Difficult 
To Determine Definitely In All Cases. — The smoothness of 
the pipe-surface, the length of time the piping has been in 
service, the size of the pipe, and the nature of the substances 
with which it may be scaled or coated internally, are the 
principal determining factors. These factors may vary 
widely, in individual cases, from working standards which are 
based upon experimental data. 

Note. — The data given herein (Tables 14 and 15) are for new pipe. 
When the pipe is very rough, or is old and rough, the actual values may 
be greater than those shown. In such cases, the resistance due to 
friction can be determined only by tests. 



10 



STEAM POWER PLANT AUXILIARIES 



[Div. 1 







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Sec. 14] PUMP CALCULATIONS 11 

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12 



STEAM POWER PLANT AUXILIARIES 



[Div. 1 



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Sec. 15] 



PUMP CALCULATIONS 



13 







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14 



STEAM POWER PLANT AUXILIARIES 



[Div. 1 



Example. — A pump (Fig. 14) draws 1050 gal. of water per min. from 
a pond and delivers it, straightaway, through 1,020 ft. of new 6-in. steel 
pipe. What pressure is required to force the water against the frictional 
resistance of the pipe? What additional pressure is required to impart 
the necessary velocity of flow in the delivery pipe? 

Solution. — By Table 1*4 the friction-head per 100 ft. of 6-in. pipe = 
9.5 ft. Hence, the friction-head for the given length of 6-in. pipe = 



Centrifugal Pump 
Priming Ejector 
B >-'- 




6- In, Delivery Pipe 

■ ■ -1020 Feet . 

I 



reservoir-. 
M 



mS^W^M^ 



-.-Suction- Pipe' 



Foot. 
Valve 





Fig. 14. — Illustrating Delivery Through Straight Run Of Pipe. 



1,020 ^ 100 X 9.5 = 96.9 ft. By For. (1), P = 0.433 L h = 0.433 X 
96.9 = 42 lb. per sq. in. 

By Table 14, the velocity = 11.9 ft. per sec. Hence, by For. (3) L hv = 
v*/2g = 11.9 2 + (2 X 32.2) = 22 ft. By For. (1) P = 0.433 L h = 
0.433 X 22 = 9.54 lb. per sq. in. This velocity-head is so small that it 
could, in practice, be neglected without appreciable error. 



16. The Heads Necessary To Overcome The Frictional 
Resistance To Water -Flow Through Fittings And Valves 

depend principally upon the ages and the sizes of the fittings 
and valves, and upon the relative smoothness of their sur- 
faces. Approximate values are given in Table 18. 

17. The Frictional Resistance Offered By The Internal 
Passages And Valves Of A Pump is very small. Often it is 
equivalent to a head loss of only 1 ft. The maximum seldom 
exceeds 3 ft. 



Sec. 18] 



PUMP CALCULATIONS 



15 



18. Table Showing Approximate Length, In Feet, Of 
Straight, Clean Wrought Iron Or Steel Pipe In Which The 
Frictional -Resistance Is Equivalent To That In The Fittings 
Listed. 



Size of pipe and fittings, in 
inches 


H 


3 A 


1 


1H 


Wi 


2 


2M 


3 


4 


5 


6 


Elbows, 90 deg. (Fig. 9) 


5 


6 


6 


8 


8 


8 


11 


15 


16 


18 


18 




2 


3 


3 


4 


4 


4 


6 


8 


9 


9 


10 






Long radius bends (Fig. 8) . . . 


2 


2 


3 


3 


3 


4 


6 


8 


9 


9 


10 


Sharp bends (Fig. 10) 


10 


12 


12 


16 


16 


16 


22 


30 


32 


36 


36 




10 


12 


12 


16 


16 


16 


22 


30 


32 


36 


36 








5 


6 


6 


8 


8 


8 


11 


15 


16 


18 


18 






Strainer or footvalve at 
entrance to suction pipe 
(Fig. 6) 


10 


12 


12 


16 


16 


16 


22 


30 


32 


36 


36 


Square kept entrance to suc- 
tion pipe (Fig. 5) 


5 


6 


6 


8 


8 


8 


11 


15 


16 


18 


18 


Funnel end entrance to suc- 
tion pipe (Fig. 7) 




































Example. — A boiler-feed pump delivers 45 gal. of water per min. 
It lifts the water, by suction, through a height (L hs , Fig. 15) of 6 ft. 
The suction piping is of 2-in. size. It extends 5 ft. below the surface of 
the water in the suction-well. It runs horizontally for a distance, L, 
of 60 ft. It makes two right-angled turns, T x and T 2 , by means of plugged 
tees, and one right-angled turn, E h by means of a 90-deg. elbow. The 
water enters the suction-pipe through an orifice, 0, which is formed by 
cutting the pipe squarely across. The delivery piping is of 1.5-in. size. 
It contains 140 ft. of straight pipe, three 90-deg. elbows, Ei, E 3 and E^ 
one globe check-valve, Vi, and one globe stop-valve, W The vertical 
height, L h d, of the discharge-lift is 35 ft. The boiler steam-pressure is 
110 lb. per sq. in. What is the total head on the pump? 

Solution. — By Table 14 the velocity in the straight runs of suction- 
piping = 4.6 ft. per sec. Also, the velocity in the straight runs of 
delivery-piping = 7.08 ft. per sec. Hence, by For. (3), L hv = v 2 /2g = 
(7.08) 2 t(2X 32.2) = 0.778 ft. = velocity-head. 

The total length of straight suction-piping = 6+5+60 = 71 ft. 
By Table 14, the friction-head due to the straight suction-piping = (71 -5- 
100) X 5.8 = 4.118 /Z. Also, the friction-head due to the straight delivery- 



16 



STEAM POWER PLANT AUXILIARIES 



[Div. 1 



piping = (140 -f 100) X 16.6 == 23.24 ft. Hence, the friction-head due 
to straight piping in the complete system = 4.118 + 23.24 = 27.358 ft. 

By Table 18, the 2-in. straight-pipe equivalent of a suction-inlet orifice 
formed by a square-cut pipe-end = 8 ft. Hence, the entrance-head, or 
friction-head due to the inlet orifice, = (8 -5- 100) X 5.8 = 0.464 ft 

By Table 18, the 2-in. straight-pipe equivalent of a sharp bend = 16 ft. 
Hence, the friction-head due to the plugged tees, T\ and T 2 , = (16 -r- 100) X 
5.8 X 2 = 1.856 ft. The 1.5-in. and 2-in. straight-pipe equivalents 
of a 90-deg. elbow = 8 ft. Hence, the friction-head due to the elbows, E\ f 
E 2 , E z , and E 4 = (8 -r- 100) X 5.8 X 4 = 1.85 J ft. The friction-head 
due to the six turns in the piping is, therefore, 1 .856 + 1.856 = 3.712 ft. 

By Table 18, the 1.5-in. straight-pipe equivalent of a globe-valve = 8 ft. 
Hence, the friction-head due to the valves V\ and Vi = (8 -*- 100) X 5.8 X 
2 = 0.928/*. 



rSucthn Inlet 



Horizontal Length of 
Suction Piping- — 



Mud Drum of 



Level Discharge \ ■ 

Valve Deck-... \ : 

:■% -l±- 

1=60-0"-- ----- -*^- 



Plugged Tees- 





fe Step- ..-\i 

"= Verves' rf\ 



Check Yalve- 



-V 



MA 



Square-Cut Orifice I 



rmmw 



Suctfon-Wdl-.w&i-x 



Fig. 15. — Pump-Piping With Large Resistance-Head. 



By Sec. 17, assume loss due to flow through passages and valves in 
chamber = 2 ft. 

By For. (4), the total friction-head = L hf T = L h f P + L h /i + L h /f + 
L hfv + L hfc = 27.358 + 0.464 + 3.712 + 0.928 + 2 = 34.462 ft. 

By transposition of For. (1) the static-head equivalent of the boiler- 
pressure = L h = 2.31 P = 2.31 X 110 = 254.1 /*. Hence, by 

For. (5), the total measured-head = L hmT = L hms + L hmd + L hmp = 
6+35 +254.1 = 295.1/*. 

By For. (6), the total head on the pump = L h r = Lhv + L h /T + 
L hmT = 0.778 + 34.462 + 295.1 = 330.34/*. 

Example. — A steam pump (Fig. 16) has a suction-lift, Lh*, of 8 ft., and 
a discharge-lift, Lhd, of 82 ft. The suction-piping is of 3-in. size. It 
contains 75 ft. of straight pipe, one long-radius bend, B, and a funneled 



Sec. 18] 



PUMP CALCULATIONS 



17 



inlet-orifice, F. The delivery-piping is of 2.5-in. size. It contains 
517 ft. of straight pipe, two 90-deg. elbows, E x and E 2 , and one globe 
valve, V. It is assumed that the head necessary to impart velocity to 
the water is (Note subjoined to Sec. 7) practically negligible. In 
practice the velocity head is, usually, practically zero. It is also assumed 
that a resistance = to 2 ft. is offered to the flow through the valves and 
passages of the pump itself. The pump discharges into an open reservor. 
It is capable of operating against a total head which is equivalent to a 
pressure of 85 lb. per sq. in. What is the maximum average-rate, 
in gals., per min., at which the pump can deliver the water through this 
system. 




J L 



Chef*** 



IX 



Td n k) 



Horizontal Length of 
' Suction Pipmg-, 



K Level of Discharge 
Valve Deck 



- ; "llhs-8ft. 




HO* 



Fig. 16. — Pump-Piping With Small Resistance-Head. 

Solution. — First, find the equivalent frictional resistances of the 
fittings in both the suction piping and the discharge piping and reduce all 
to the basis of 3-in. piping as explained below: By Table 18, the 3^in. 
straight-pipe equivalent of the long-radius bend, B, =8 ft. Also, the 
straight-pipe equivalent of the funneled inlet-orifice, F, — 0.0 ft. Hence, 
the frictional resistance in the complete 3 in. suction-piping is that which 
would occur in a straight run of 75 + 8 = 83 ft. of 3-in. pipe. 

By Table 18, the 2.5-in. straight-pipe equivalent of the two 90-deg. elbows, 
Ei and E 2 , = (11 X 2) = 22 ft. Also, the straight-pipe equivalent of the 
globe-valve, V, = 11 ft. Hence, the frictional resistance in the complete 
2.5-in. delivery-piping is that which would occur in a straight run of (517 + 
22 + 11+2) = 552 ft. of 2.5-in. pipe. 

Now by comparing the Friction-Head values for "2%-Inch Pipe" 
and for ic 3-Inch Pipe" from Table 14, it will be found that, on the average, 
23^-in. pipe offers 2.4 times as much frictional resistance for the same 
flow, in gallons per minute, as does 3-in. pipe. Hence, the frictional 
resistance in the entire piping system is equivalent to that which would occur 
in a straight run of: 83 + (552 X 2.4) = 1408 ft. of 3-in. pipe. 
2 



18 STEAM POWER PLANT AUXILIARIES [Div. 1 

By For. (5), the total measured head, L hmT = L m hs + L mhd = 
8 + 82 = 90 ft. 

By transposition of For. (1), the total static head developed by the pump 
which is equivalent to a pressure of 85 lb. per sq. in. = L h = 2.31P = 
2.31 X 85 =. 196.35 ft. 

Hence, the head which remains or which is available for overcoming the 
frictional resistance of the entire pumping system, that is, the frictional 
resistance or head of 1408 ft. of 3-in. pipe = 196.35 - 90 = 106.35 ft. 
Stating this in friction head per 100 ft. of straight pipe: (106.35 -f- 1408) 
X 100 = 7.55 ft. friction head per 100-ft. length of 3-in. pipe. 

By Table 14, the flow corresponding to a friction-head of 7.72 ft. per 
100-ft. length of 3-in. pipe = 150 gal. per min. Hence, the flow corre- 
sponding to a friction head of 7.55 ft. per 100-ft. length of 3-in. pipe is, 
approximately: (7.55 X 150) -s- 7.72 = 146.7 gal. per min. = the maxi- 
mum average rate of delivery. 

19. The Proper Sizes For The Suction Or Discharge Pipe 
Of Any Pump may be computed with the following formula, 
the derivation of which is given below : 

(7) dj = 4.95 J^* (inches) 

Wherein, d, = actual internal diameter of suction — or dis- 
charge — pipe, in inches. V gm = amount of water to be pumped, 
in gallons per minute. v m = average velocity of flow, in 
feet per minute. 

Note. — Transposing the above there results: 

(8) v m = — '-^— 9Jn . (feet per min.) 

a j 

(9) V gm = ^r = 0.004 dH m (gallons per min.) 

Derivation. — Since, V e f, the amount of water to be pumped in cubic 
feet per minute = (A /, the cross-sectional area of the suction or discharge 
pipe, in square feet) X (v m , the allowable velocity of flow, in feet per minute), 
it follows that: 

(10) Vef = AfV m = | ( y 2 ) V™ = -57^ ( Cir - ft - P er min -) 

Also, since 1 cu. ft. = 7.48 gal: 

(cu. ft. per min.) 



(11) 


-it " am 

7c/ = 7^8 


Now, 


equating (10) and (11): 


(12) 


V am irdi 2 Vm 

7.48 576 



Sec. 20] 



PUMP CALCULATIONS 



19 



Then, solving for d t : 
(13) dt 



4.95 



& 



(inches) 



Example. — A simple direct-acting steam pump is required to deliver 
800 gals, of water per min. What should be the diameter of the suction 
pipe if the allowable flow velocity in it (See Sec. 52) is 200 ft. per min.? 
What should be the diameter of the discharge pipe if the allowable flow 
velocity in i is 400 ft. per min.? Solution. — For he suction pipe 
by For. (7), di = 4:.95VV gm /v m = 4.95 V800 h- 200 = 9.9 in., or, practi- 
cally, a 10-in. internal-diameter pipe. For the discharge pipe: di = 
4.95\/800/400 = 6.98 in. or, practically, a 7-in. internal-diameter pipe. 

20. The Displacement Of A 
Reciprocating Pump is the vol- 
ume of space (Fig. 17) swept 
through by the piston or plunger 
in a definite interval of time. 
Assuming the pump-cylinder to 
be full of water at the beginning 
of each stroke, the displacement 
is equal to the volume of water 
which is driven out of the cylinder 
during the given time-interval. 

Note. — The displacement of a pump 
may be expressed in cubic feet, pounds or 
gallons per minute. 



Discharge 
Outlet-. 



Piston-. 



21. The Displacement Of Any 
Piston Or Plunger Pump Per 

Minute may be found by the fol- 
lowing formulae: 

(14) V cf - LANs 




Suction 
Inlet- 



Fig. 17. — Showing Volume Of 
Space Swept Through In One 
Stroke Of Piston. 



(15) 
(16) 



1728 
W. = LAN s Di 
LANs 



V = 



231 



(cubic feet per min.) 
(pounds per min) 
(gallons per min) 



Wherein V c f = displacement in cubic feet per minute. W» 
= displacement in pounds per minute. V gm = displacement 
in gallons per minute. L = length of stroke in inches. A = 
effective area of piston or plunger, in square inches. N 8 = 
number of strokes per minute. Di = density of liquid to be 
pumped in pounds per cubic inch. 



20 



STEAM POWER PLANT AUXILIARIES 



[Div. 1 



Note. — The Effective Plunger Or Piston Area of an outside- 
end-packed plunger (Fig. 18) in a direct-acting steam-pump is the 
cross-sectional area of the plunger. Of a center-packed (Fig. 19) or 
inside-packed (Fig. 20) plunger, or of a piston (Fig. 17), it is the cross- 



■Pisfon Plunger-Root 
Cradle-Rods-. 



Discharge Outlet 

Plunger-.. _ 
Cross head "v 

Packing ', 
Gland- -. 




***"*- suction Inlet- 

^mwWx\\\\\\\\\^ 



^^W> 



Fig. 18. 



-Water-End Of Direct-Acting Steam-Pump With Outside End-Packed 
Plungers. 



sectional area of the plunger or piston minus one-half the cross-sectional 
area of the piston- or plunger-rod. 

Example.— What is the displacement, in cubic feet per minute, of an 
outside center-packed duplex pump (Fig. 19), if: the plunger-diameter 
is 18 in., the plunger-rod diameter is 3 in., the length of stroke is 24 in., 



Gland-, 




\^\<\^\\\X\\\\V\\\\\V 



Fig. 19. — Water-End Of Direct-Acting Steam- 
Pump With Outside Center-Packed Plungers. 



Fig. 20. — Pump-Plunger Inside- 
Packed With Fibrous Rings. 



and each of the 2 plungers (this being a duplex pump) makes 50 strokes 
per min.? Solution. — By preceding Note, the effective plunger area = 
(18 2 X 0.7854) - (3- X 0.7854 + 2) = 250.8 sq. in. Now, substitute 
in For. (14): V cf = LAN S /1728 = 24 X 250.8 X 50 X 2 -r 1728 = 
348.5 cu. ft. per min. 



Sec. 22] 



PUMP CALCULATIONS 



21 



Discharge Outlet 

Valves V.anol 
ViinActof 

Closing. 

Suction 
.--Inlet 



.-Discharge Valves 



Suction 
Yctlves 



22. Pump-Slip is the return of water, or other liquid, 
through the valves of a pump while the valves (Fig. 21) are in 
the act of closing. It may also occur by leakage past 
the piston or plunger from the discharge-end to the 
suction-end while the pump is making a stroke. It is, there- 
fore, the difference between the displacement or theoretical 
discharge of a pump and the actual discharge. It is commonly 
expressed as a percentage of the displacement. 

Note. — Average Values Of Pump- 
Slip, for good pumps, range from 3 to 
5 per cent. The slip of a new pump 
seldom exceeds 2 per cent. Where 
conditions are adverse, the slip may 
be as great as 10 or 15 per cent. For 
pumps which handle large volumes of 
water, slips as low as % per cent, 
have been recorded. 

Note. — Pump-Slip May Be Nega- 
tive. That is, the actual discharge 
may be greater than the theoretical 
discharge. This may occur if the 
suction-lift is very low, and the 
suction- and discharge-lines run hori- 
zontally for considerable distances. 

The momentum of the moving column (Sec. 65) may then cause the 
suction water to surge into the cylinder with such force as to produce a 
considerable leakage through the discharge valves at the suction end. 

23. Very High Piston-Speed May Cause Excessive Pump- 
Slip. — When the piston reaches the end of a stroke, a space of 
time must elapse while the open valves (Fig. 21) are descending 
and making firm contact with their seats. But, during this 
interval, the piston starts on the opposite stroke. Some of the 
water that was discharged during the preceding stroke then 
flows in behind the piston through the imperfectly seated dis- 
charge valves. Admission of a full cylinder of water through 
the suction valves is thus prevented. Coincidentally, some 
of the water ahead of the piston slips by the suction valves and 
passes back into the suction chamber. 

24. The Percentage Of Pump -Slip May be computed by the 
following formula: 

v 100(7 C/ - V a ) 




Piston- 

^\\\\\\\\\\r\\\\\\\\\\\\v 

Fig. 21. — How Pump-Slip Occurs. 



(17) 



Vcf 



(per cent.) 



22 STEAM POWER PLANT AUXILIARIES [Div. 1 

Wherein X = per cent, of slip. V c f = displacement in cubic 
feet per minute. V a = actual discharge in cubic feet per minute. 
Example. — The displacement of a pump is 386.85 cu. ft. per min. 
The pump delivers 372.4 cu. ft. of water per min. What is the slip? 
Solution.— By For. (17), X = [100 (F c/ - 7«>] -^ V cf = [100 X 
(386.85 - 372.4)] ^ 386.85 = 3.74 per cent. 

25. The Volumetric Efficiency Of A Pump is the ratio of the 
volume of water actually delivered by the pump to the dis- 
placement of the pump. It may be computed by the following 
formula: 

(18) E r = ^~ (per cent.) 

V cf 

Wherein E v = the volumetric efficiency, in per cent. V a = 
the actual discharge, in cubic feet per minute. V c / = the 
theoretical discharge, or the displacement, in cubic feet per 
minute. 

Note. — Volumetric efficiency and pump-slip are closely related. 
Thus, pump-slip = 1- volumetric efficiency. Pump-slip may vary from 
0.5 per cent, to 15 per cent. Hence, the volumetric efficiency may 
correspondingly vary from 99.5 per cent, to about 85 per cent. Pump- 
slip exceeding 2 per cent, would indicate either unfavorable operating 
conditions, defective design, or a worn-out condition of the pump. 

26. The Discharge Of A Piston Or Plunger Pump may be 

approximately computed by the following formula : 

(19) Va = j~ E vd = |gg 35 (cubic feet per mm.) 

Wherein V a — approximate discharge capacity, in cubic feet 
per minute. d p = diameter of piston or plunger, in inches. 
L T = the effective piston or plunger travel, in feet per minute. 
E vd = the volumetric efficiency, expressed decimally. 

Example. — The plunger diameter in a direct acting duplex steam pump 
is 6 in. The stroke is 24 in. Each plunger makes 35 strokes per min. 
What is the discharge when the volumetric efficiency is 92 per cent.? 
Solution. — The total number of strokes per minute, this being a duplex 
pump, = 2 X 35 = 70. By For. (19), V a = d p 2 L T E vd / 183.35 = [6 2 X 
(70 X 24 + 12) X 0.92] -r- 183.35 = 25.3 cu. ft. per min. 

27. The Requisite Diameter For The Water-End Of A Pump 
Plunger Or Piston, when the rates of discharge and plunger or 



Sec. 28] PUMP CALCULATION 23 

piston travel are given, may be found by the following 
formula : 

(20) d p = ^^~~ (inches) 

Wherein d p = the diameter of plunger or piston, in inches. 
V a = the actual discharge, in cubic feet per minute. L T — 
the effective travel of the plunger or piston, in feet per min- 
ute. E vd = the volumetric efficiency of the pump, expressed 
decimally. 

Example. — A single direct acting steam pump is required to discharge 
141 gal. of water per min. while running 90 ft. of plunger travel per min. 
If the assumed volumetric efficiency of the pump is 97 per cent., what 
should be the diameter of the plunger? Solution. — A gallon contains 
231 cu. in. By For. (20), d p = Vl8Z.35V a /L T E vd = 
V183.35 X (141 X 231 + 1728) -=- (90 X 0.97) = 6.3 in. 

28. The Requisite Steam-Piston Diameter For A Direct- 
Acting Steam Pump may be found by the following formula : 



(21) d s = yj—p — - (inches) 

Wherein d s = diameter of steam-piston, in inches. A = 
area of water-piston or plunger, in square inches. P w = total 
head-pressure, in pounds per square inch. P s = steam pres- 
sure, in pounds per square inch. The mechanical efficiency 
(Sec. 41) of the pump is herein assumed as 70 per cent. 

Example. — The requisite diameter of water-piston for a direct-acting 
steam pump is found to be 8 in. The total head-pressure is 200 lb. per 
sq. in. The available steam-pressure is 80 lb. per sq. in. What should be 
the steam-piston diameter? Solution. — By For. (21) d, = \Zl.SAP w /P s 
= Vl.8 X 8 2 X 0.7854 X 200 + 80 = 15 in. 

Note. — The Plunger- Or Water-Piston-Size For A Duplex 
Pump Is Computed On The Basis of one-half the total quantity of water 
to be delivered, and upon the rate of travel of one piston. 

Note. — The Piston-Speed Of A Direct-Acting Steam-Pump should 
be gaged according to the size of the pump. In large- and medium-sized 
pumps for general service, it should not exceed about 100 ft. per min. 
In small pumps, with strokes of from about 3 to 9 in., the piston travel 
should range from about 40 to 75 ft. per min. 



24 STEAM POWER PLANT AUXILIARIES [Div. 1 

29. To Compute The Average Velocity Of Flow Through 
The Discharge Pipe Of Any Reciprocating Pump, the following 
formula may be used. Slip is disregarded. 

(22) v m = * 2 T (feet per minute) 

Wherein, v m = the average velocity of flow through the dis- 
charge pipe, in feet per minute. d p = diameter of water piston, 
in inches, di = actual internal diameter of discharge-pipe, in 
inches. L T = effective piston travel, in feet per minute ; for a 
double-acting pump, L T = feet which the piston travels in a 
minute; for a single-acting pump, L T = {feet which the piston 
travels in a minute) -^2. 

Example. — The water-piston diameter in a direct-acting (double- 
acting) steam pump is 3^ in. The discharge-pipe internal diam. is 
\}4, in. The piston travel is 100 ft. per min. What is the average 
velocity of water flow in the discharge pipe? Solution. — By For. (22): 
v m = d 2 pLT/di 2 p= 3.5 X 3.5 X 100 -r (1.5 X 1.5) = 544 ft. per min. 

30. The Net Work Of A Pump is the quantity of work 
which is theoretically necessary to elevate the water or other 
liquid from the suction-level to the discharge-level. That is, 
it is the work performed in overcoming the total measured 
head, L hmT For. (5). 

31. The Net Work Performed By A Pump May Be Com- 
puted by the following formula : 

(23) W u = WL hmT (foot-pounds) 

Wherein W u = net work in foot-pounds. W = weight of 
water or other liquid pumped, in pounds. L hmT = the total 
measured head (Sec. 11), in feet = vertical height, in feet, 
from level of suction supply to discharge level. 

Example. — A pump lifts 14,620 lb. of water from a pond and de- 
livers it to a reservoir. The vertical distance between the suction- and 
discharge-levels is 41 ft. What is the net pump-work? Solution. — By 
For. (23), W u = WL hm T = 14,620 X 41 = 599,420 ft. lb. 

32. The Actual Work Of A Pump includes, in addition to the 
net work (Sec. 30), all of the work performed in overcoming 
frictional resistances and in imparting velocity to the liquid. 



Sec. 33] 



PUMP CALCULATIONS 



25 



• Aolmi55ion Port- 
Steam Line—- Closure 

Admission Admission 

■Port Opens ..-Line 



■Exhaust 

Port 
Closure 



Back- 
Pressure 
Line-. 



Exhaust 

Port 

Opens—- 



Exhaust 
Line — 



Atmospheric Line—' 

Fig. 22. — Indicator Diagram From Steam-Cyl- 
inder Of Direct-Acting Steam-Pump. 



The frictional resistances include, besides the water-friction 
in the suction- and discharge-pipes and in the pump passages, 
the mechanical friction between the moving parts of the pump- 
ing mechanism. 

33. The Rate At Which A Pump Does Work May Be 
Expressed In Terms Of Horse Power. — The total horse 
power developed in the 
steam-cylinders of steam- 
pumps may be computed 
from indicator diagrams 
(Fig. 22) taken from the 
steam-cylinders. The total 
horse power developed in 
the water-cylinders of re- 
ciprocating pumps of all 
types may be computed from indicator diagrams (Fig. 23) 
taken from the water-cylinders. 

Explanation. — In the pump diagram (Fig. 23) the total height, 
LhT, indicates, to the scale of the diagram, the total head, For. (6), on 
the pump; this is called the indicated head. The heights Lhms and Lhmd 
indicate, respectively, the measured suction head, Lhms of For. (5), and 
the measured delivery head, L hm d, For. (5). The heights s and d indicate, 
respectively, the friction heads on the suction and delivery sides of the 

pump. That is, s + d indicates 
the total friction head, L h jT of For. 
(4). The sum L hms + L hmd + d 
comprises the useful head on the 
pump. The velocity head is 
herein considered as being so 
small that it may be neglected. 
All of these heads are expressed 
(Sec. 5) in pounds per square inch. 



Pressure Variation 
Due to Trembling' of 
Valves, Caused 



Total 
Head-- 



by Spring Re- 



Friction Head 



action and 
Inertia of . rta 
Water- " 



on delivery Side ; • fjeacl 



.-Measured 
;' Delivery 



~ 



L hT '' L hmd 
Atmospheric Line'* 



.LjhHT^ 



Measured 

■Suction 

Head 



Friction Head oh--' 
Suction Side 



Absolute Zero- Pressure 



34. The Hydraulic Or 
Water Horse Power Devel- 

riG. 23. — Indicator Diagram From 
Water-Cylinder Of Reciprocating Pump. Oped By A Pump is the USe- 
( Velocity head is hereon neglected.) f^Q horge power developed in 

the pump cylinder as computed upon a basis which comprises 
the actual weight of water discharged and the total useful head. 
It may be expressed by the following formula : 

Wi m L hu 



(24) 



uhp 



33000 



(horse power) 



26 



STEAM POWER PLANT AUXILIARIES 



[Div. 1 



Wherein T? U h P = the theoretical hydraulic horse power. 
Wz TO = the weight of liquid pumped, in pounds per minute. 
Lhu = the total useful head, in feet, = the head, in feet, corre- 
sponding to the gage-pressure (Fig. 24) at the pump discharge 
nozzle + the head due to the height of the discharge nozzle 
above the level of the source of suction supply. 



(Pressure Gage ;$atz Valve 
.-Discharge 
• Nozzle 




w — -- 

: Centrifugal. 
Pump • 

-Suction Pipe- 

.'-^-foot-Valve 






Fig. 24. — Pump Showing 40 
Lb. Per Sq. In. Gage Pressure 
At Discharge Nozzle. 

power, as computed 
diagram (Fig. 23). 



Example. — A direct-acting steam-pump 
moves 4,160 lb. of water per min. against 
a total useful head of 36 ft. What is the 
net horse power developed? Solution. — 
By For. (24), V uhp = W^L^/33000 = 4160 
X 36 -r 33000 = 4.5 h.p. 

35. The Indicated Efficiency Of A 
Reciprocating Pump is the ratio, ex- 
pressed as a per cent., of the net useful 
horse power (Sec. 34) to the horse 
power computed (Sec. 39) from the 
pump indicator diagram (Fig. 23). It 
may be expressed by formula: 

(25) E, = i^^ (per cent.) 

" whp 

Wherein E»- = the indicated efficiency, 
in per cent. *P u k v = the useful horse 
power (Sec. 34). V whp = the horse 
(Sec. 39) from the pump indicator 



Note. — The indicated efficiency of a pump is a criterion of the sum 
total of hydraulic losses, or of the losses occurring solely in the water-end. 

Example. — A reciprocating plunger pump moves 5910 lb. of water 
against a total useful head of 61 ft. The hydraulic horse power 
developed, as computed from an indicator diagram, is 12.14. What is 
the indicated efficiency? Solution. — By For. (24), the useful horse power 
= ~Puh P = Wi ro L Att /33,000 = 5910 X 61 -^ 33,000 = 10.93 h.p. By For. 
(25), the indicated efficiency = E* = 100 'P u h P /'Pwhp = 100 X 10.93 -S- 
12.14 = 90 per cent. 

36. The Hydraulic Losses Of A Pump are defined as those 
losses in hydraulic pressure (or head) which occur in the 
suction pipe and in the pump itself. They comprise pressure 
equivalents of the losses in head due to: (1) The Passage 



Sec. 37] 



PUMP CALCULATION Si 



27 



Of The Water From The Well Or Other Supply Source, 
Through The Suction Pipe And Pump, To The Point 
Where The Discharge Gage {D, Fig. 25) Is Connected; 
these consist of : (a) suction-pipe entrance loss, (b) suction-pipe 
and pump velocity loss, (c) suction-pipe friction loss, (d) losses 
in suction-pipe bends and connections, (e) friction loss in 
passing through pump suc- 
tion valves, (f) friction loss 
in passing through pump 
discharge valves. (2) The 
Pressure Necessary To 
Overcome The Reaction 
Of The Springs Of The 
Discharge Valves. The 
pressure lost due to losses 
under (1) and (2) will each 
be equivalent to about % 
per cent, of the total dis- 
charge pressure, giving a 
total hydraulic loss of 
about 1 per cent. 



Gage For Indicating! 
Pressures Above Atmos 
pher/'c Pressure 



Measured Head Between 
Points Of Attachment Of I 
Gorges-...^ 




Hand holes 

Exhaust \ F or Access 

Connection 1 To Valves-' 

Gage For Indicating Pressures Above 1 
Or Below Atmospheric Pressure 

Fig. 25. — Duplex Fire-Pump With Dis- 
charge And Suction Gages Attached. De- 
signed To Run From 150 To 250 Ft. Of Piston 
Travel Per Min. Steam Cylinders, 14-ln. 
Diameter. Water Cylinders, 8.5-In. Diameter. 
Stroke, 12-ln. 



Note. — In commercial pump 
tests and computations, it is, as 
above indicated, ordinarily un- 
derstood that the hydraulic 

losses in the suction pipe are to be included with the losses in the pump 
itself. From a theoretical standpoint, this is incorrect. But the pump 
manufacturers accept this practice because it simplifies testing and 
guarantees. In any case, the true suction-pipe losses are very small 
and will be practically the same for all pumps which are doing the same 
work. On the other hand, the discharge-pipe losses are never included 
in the hydraulic losses of a pump. 

37. The Hydraulic Efficiency Of A Pump may be expressed 
as a percentage by the formula: 

P + the hydraulic losses 

Wherein: E& = the hydraulic efficiency, in per cent. 

P = [pressure as read on discharge gage (Z>, Fig. 25) in 

lb. per sq. in., when pump is delivering the quantity of water 

at which it is desired to determine EJ + [0.433 X (distance in 



28 



STEAM POWER PLANT AUXILIARIES 



[Div. 1 



feet from discharge gage to surface of water in well)]. The 
hydraulic losses are as enumerated above; they may be 
obtained as explained in the following note. 

Note. — The Necessary Data To Determine The Hydraulic 
Efficiency of a given pump may be secured in the following manner: 
An indicator is attached to the water cylinder and the pump driven at 
such a speed and the discharge valve is so throttled that the pump will 
deliver that quantity of water at which the hydraulic efficiency is de- 
sired. The discharge valve must be located on the discharge side of 
gage D and some distance away from it. 
Hydraulic indicator cards (Fig. 26) are then 
taken, and at the instant the card is taken the 
discharge gage (D, Fig. 25) is read. Compute P 
as above indicated, and lay off this pressure (line 
AB, Fig. 26) to the scale of the indicator card, 
measuring downward from the top of the indi- 
cator card as shown in Fig. 26. The remainder 
of the distance, BC, is, to the scale of the indi- 
cator card, the hydraulic losses, that is the pressure 
required to overcome the losses. 




Total Hydraulic Losses 



Fig. 26.— Indicator Card 
Taken On Water End Of 
Steam Pump, Showing 
Total Hydraulic Losses. 



38. The Usual Practice In Determining The Load On A 
Pump is to attach a pressure-gage to the discharge-pipe, 
D (Fig. 25), and to the intake-pipe, S (Fig. 25), a gage which 
indicates both vacua, or pressures below atmospheric, and 
pressures above atmospheric. Then, if the pump lifts the 
water, the suction-gage, S, will indicate a pressure less than 
atmospheric. But if the water flows, under a head, to the 
intake of the pump, the intake-gage, S, will indicate a pressure 
greater than atmospheric. If $ indicates a pressure less than 
atmospheric, the net gage-pressure is found by adding, to the 
pressure per square inch shown by the discharge-gage D, the 
pressure per square inch which corresponds to the vacuum, 
in inches, shown by the intake gage, S. If $ indicates a 
pressure above atmospheric, the net gage-pressure is then 
found by subtracting, from the pressure per square inch 
shown by the discharge-gage, D, the pressure per square inch 
above atmospheric, which is shown by the intake-gage, S. 
The load on the pump, in pounds per square inch, is then 
equal to the net gage-pressure plus the pressure which is due 
to the hydrostatic head, L hm (Fig, 25), between the points of 
attachment of the gages. 



Sec. 39] PUMP CALCULATIONS 29 

Example. — If the discharge-gage, D, (Fig. 25) shows 41 lb. per sq. 
in., and the intake-gage, S, shows a vacuum corresponding to a reduc- 
tion of 3 lb. per sq. in. below normal atmospheric pressure, then the 
net gage-pressure is 41 +3 = 44 lb. per sq. in. And if the vertical 
height, Lhm, is 5 ft., then the pressure against which the pump works = 
44 + (5 X 0.433) = 46.1 lb. per sq. in. But if the intake-gage shows a 
pressure of 3 lb. above normal atmospheric pressure, then the pressure 
against which the pump works = [41 — 3] + [5 X 0.433] = 40.1 lb. per 
sq. in. 

Note. — Conversion Of A Vacuum Reading In Inches Of Mer- 
cury To Terms Of Pounds Per Square Inch may be done by multi- 
plying the reading in inches by 0.4914, or, in practice by 0.49. 

Example. — If the-intake-gage, S, (Fig. 25), shows a vacuum of 5 in., 
the difference between normal atmospheric pressure and the pressure in the 
intake-pipe = 5 X 0.49 = 2.25 lb. per sq. in. 

39. The Actual Or Indicated Hydraulic- or Water-Horse- 
Power Developed By A Pump is the total horsepower devel- 
oped in the pump cylinder, as computed upon a basis of the 
mean pressure throughout the discharge stroke of the pump 
plunger. The mean pressure is obtained from an indicator 
diagram (Fig. 23) taken during a double stroke of the pump 
plunger. The hydraulic horsepower computed upon this 
basis includes the power expended in overcoming all resistance 
due to water-friction from the inlet orifice of the suction pipe to 
the outlet orifice of the discharge pipe. The indicated hydraulic 
horsepower may be expressed by the following formula : 

(27) T? whp = (H.P.) I = P ^ f ^ s (horsepower) 

Wherein P whp = the actual hydraulic horsepower. P = the 
load on the pump, Sec. 38, in pounds per square inch, which 
may be computed from the indicator diagram. L f = the 
length of the stroke, in feet; A = the area of the plunger, in 
square inches. N a = the number of strokes per minute. 

40. The Total Driving Horse Power Developed By A 
Steam Pump Or Delivered To A Power Pump includes the 
actual hydraulic horsepower (Sec. 39) plus the horsepower 
required to overcome the mechanical or metal-to-metal friction 
in the complete pumping mechanism. In the case of a steam- 
pump, the total driving horse-power will correspond to the 
indicated horse-power, as computed with the aid of indicator 
diagrams (See the author's Steam Engines) which is developed 



30 STEAM POWER PLANT AUXILIARIES [Div. 1 

in the steam cylinder or cylinders. In the case of a power 
plunger-pump, or of a centrifugal pump, the total driving 
horsepower will lbe the horsepower delivered by belt-trans- 
mission, gear-transmission, or by direct motor-connection, to 
the pump pulley, driving-shaft or spindle. 

41. The Mechanical Efficiency Of A Reciprocating Pump 
is the ratio, expressed as a per cent., of the indicated hydraulic 
horsepower (Sec. 39) to the driving horsepower. The hydrau- 
lic horsepower may be computed from (Fig. 23) a pump indica- 
tor diagram. The driving horsepower of a steam-driven 
pump may be computed from (Fig. 22) a steam indicator 
diagram. In the case of a power pump (Fig. 24) the driving 
horsepower is the total horsepower delivered to the pump by 
belt, gearing, or direct shaft-connection. The mechanical 
efficiency may be expressed by the following formula : 

(28) E m = 100P ^ = (per cent.) 

JTbhp 

Wherein E m = the mechanical efficiency, in per cent. P W h P 
= the hydraulic horsepower, as computed from the pump 
indicator diagram. Pi hp = the driving horsepower. 

Note. — The mechanical efficiency of a pump is a criterion of the loss 
due to mechanical friction in the mechanism which transmits the driving 
power to the water end of the pump. The higher the mechanical effi- 
ciency the less the mechanical losses in the pump. 

Example. — The hydraulic horsepower of direct-acting steam pump, as 
computed from a pump-indicator diagram is 42.3. The driving horse- 
power, as computed from a steam-indicator diagram, is 49.76. What is 
the mechanical efficiency. Solution. — By For. (28): E m = 100 P W h P / 
Pbhp = 100 X 42.3 + 49.76 = 85 per cent 

Note. — The Maximum Mechanical Efficiency Obtained With 
Direct-Acting Steam Pumps is about 80 per cent. This efficiency may 
be had with very large pumps. The efficiencies diminish with the sizes 
of the pumps. Very small pumps may give an efficiency of only 50 per 
cent., or even less. 

42. The Total Efficiency Of A Pump Is The Product Of The 
Volumetric, Hydraulic, And Mechanical Efficiencies. It is 

the efficiency which is, ordinarily, specified by the manufacturer 
of the pump. It is a criterion of the pump's overall economy 
in the use of power. It may be expressed by formula : 

(29) Be = ^ q q (per cent.) 



Sec. 43] 



PUMP CALCULATIONS 



31 



Wherein E* = the total efficiency, in per cent. E v = the 
volumetric efficiency, in per cent. E^ = the hydraulic effi- 
ciency, in per cent. E m = the mechanical efficiency, in per 
cent. 

Note. — For a steam-driven pump, the total efficiency recognizes all 
losses — steam, mechanical and hydraulic — from the steam cylinder 
to the water-discharge pipe. For a power-driven pump, the total effi- 
ciency recognizes only the mechanical and hydraulic losses from the 
driven pulley, gear or shaft to the water-discharge pipe. 

43. Total-Efficiency Values For Different Pumps may vary 
widely with the condition and the design of the pump. Cen- 
trifugal pumps may show total efficiencies thus:-100 gal. per 
min., 40 per cent.; 200 gal., 50 per cent.; 300 gal., 60 per cent.; 
400 gal., 65 per cent., 600 gal., 70 per cent.; 800 gal., 85 per 
cent.; 100 gal., 75 per cent.; 1500 gal., 78 per cent. The 
efficiency of a centrifugal pump is also determined largely by 
its speed and capacity. Hence it is always advisable, when 
specific data are required, to obtain guarantees from the 
manufacturers. The total efficiency of a belt- or gear-driven 
power pump may range from about 50 to 80 per cent. 

44. Table Showing Approximate Total Efficiencies Of 
Steam Pumps In Good Condition (Peele's Mining Engineers' 
Handbook). 





Total efficiency in per cent. 


Stroke 


Non- 


Compound 


Compound 


Triple- 




condensing 


non- 
condensing 


condensing 


expansion 
condensing 


4 


21 








6 


26 


26 






8 


30 


30 






10 


34 


34 


41 


50 


12 


37 


37 


45 


54 


15 


40 


40 


48 


58 


18 


43 


43 


52 


62 


24 


47 


47 


55 


66 


36 




50 


59 


70 


48 






63 


74 



32 STEAM POWER PLANT AUXILIARIES [Div. 1 

45. The Horsepower Required For Pumping may be com- 
puted by the following formula : 

(30) J? bhp = T (horsepower) 

Wherein P& Ap = the horsepower input required to drive a 
pump against the maximum total head ; for a steam pump it is 
the indicated steam horsepower required for the steam end, 
for a power-pump it is the horsepower input required at the 
driving pulley, gear or shaft. W Zm = the weight of water to be 
pumped, in pounds per minute. L h T = the total head on the 
pump, in feet. E* = the total efficiency of the pump, in per 
cent., as defined in Sec 42. 

Example. — It is required to pump 1,205 gal. of water per min. against 
a total head of 450 ft. The total efficiency of the pump which will be 
used is 64 per cent. What horsepower must be supplied to operate the 
pump? Solution. — Since 1 gal. of water weighs 8.3 lb., 1,205 gals, will 
weigh: 1,205 X 8.3 = 10,000 lb. By For. (30); P«* p = W, TO W330E* = 
(10,000 X 450) -=- (330 X 64) = 213 h.p. 

Example. — It is required to pump 10,000 lb. of water per min. against 
a total head of 450 ft. Assuming volumetric and hydraulic efficiencies 
of 98 per cent, each, and a mechanical efficiency of 80 per cent., what 
horsepower must be supplied? 

Solution.— By For. (29), the total efficiency = E t = E,E A E m /10,000 = 
98 X 98 X 80 -r 10,000= 77 per cent. By For. (30), the required horse- 
power = P bhp = W* JW330 E t = (10,000 X 450) -f- (330 X 77) = 177 
h.p. 

46. The Duty Of A Steam Pump is the ratio of the work done 
by the pump to the quantity of coal, steam or heat consumed 
in doing the work. 

47. The Duty Of A Steam Pump On A Basis Of Coal Con- 
sumption may be found by the following formula : 

(31) D c = 100 ™ LhT (ft. lb. per 100 lb. coal) 

Wherein D c = duty, foot pounds, per 100 lb. of coal. W = 
weight of liquid pumped, in pounds. L hT = total head on 
pump in feet. W c = weight of coal consumed in pounds. 

Example. — A steam pump raises 12,900,000 lb. of water against a 
total head (Sec. 12) of 60 feet. The steam supplied to the pump, while 
doing this work, requires the combustion of 2,500 lb. of coal. What is the 
duty? 



Sec. 48] PUMP CALCULATIONS 33 

Solution.— By For. (31), D c = 100 WL hT /W c = 100 X 12,900,000 X 
60 4- 2,500 = 30,960,000 ft. lb. per 100 lb. of coal. 

Note. — Pump-Duty Computed On A Basis Of Coal Consumption 
is of practical use in comparing the merits of two or more steam pumps 
only when the same quality of coal is used in testing all of the pumps. 

48. The Duty Of A Steam Pump On A Basis Of Steam 
Consumption may be computed by the following formula : 

(32) D s = 1000 ™ LhT (ft. lb. per 1000 lb. steam) 

W s 

Wherein D s = duty, in foot pounds per 1,000 lb. of dry steam. 
W = weight of water pumped, in pounds. L hT = total 
head on pump in feet. W s = weight of steam consumed, 
in pounds. 

Example. — A steam-pump raises 8,765,000 lb. of water against a 
total head of 125 ft. The steam-consumption is 8,315 lb. What is the 
duty? 

Solution.— By For. (32) D s = 1,000 WL hT /W s = 1,000 X 8,765,000 X 
125 -7- 8,315 = 131,764,883 ft. lb. per 1,000 lb. of dry steam. 

Note. — Pump-Duty Computed On A Basis Of Steam-Consumption 
may have only an approximate value. This may be due to the difficulty 
of determining the exact weight of dry steam used. It may also be due to 
variations of steam pressure. A given weight of high-pressure steam 
will do more work in the cylinder than the same weight of compara- 
tively low pressure steam. 

, 49. The Duty Of A Steam Pump On A Basis Of The Quan- 
tity Of Heat Consumed may be computed by the following 
formula : 

, QQ . _ l,000,000(P d + Pi + P D )AL f N s 1,000,000 WL hT 

{66) D h = g = jj. 

(ft. lb. per 1,000,000 B.t.u.) 
Wherein D/> = duty, in foot pounds, per 1,000,000 B.t.u. 
P d = discharge pressure, in pounds per square inch, as indicated 
by a gage in the discharge pipe. Pi = intake pressure, in 
pounds per square inch, as measured from atmospheric pres- 
sure (Sec. 38) by a gage in the intake pipe — to be added if 
negative and to be subtracted if positive. P D = pressure, in 
pounds per square inch, due to hydrostatic head between 
points of attachment of pressure gages. A = effective area of 
plunger, in square inches. Lj = length of stroke, in feet. N s 



34 STEAM POWER PLANT AUXILIARIES [Div. 1 

= total number of strokes H = total quantity of heat 
consumed, in British thermal units, as determined by steam 
consumption test; see the author's Steam Engines. 

Note. — Pump-Duty Computed On A Basis Of Heat-Consumption 
is more nearly exact than computations (Sec. 49) on bases of coal- or 
steam-consumption. Since the determining factor is the actual quan- 
tity of heat energy expended in the steam-cylinder, pump-duty figured 
on this basis provides a true criterion of the comparative working effi- 
ciencies of two or more different pumps. This method has been recom- 
mended by the A. S. M. E. 

Example. — A duplex steam-pump has inside-packed plungers of 20- 
in. diameter and 15-in. stroke. The plunger-rods are of 3-in. diameter. 
The total heat in the steam supplied to this pump, during a duty trial, 
was 17,642,400 B.t.u. The pump made, during the trial, 37,264 strokes. 
The average discharge-pressure, as indicated by a gage in the discharge 
pipe, was 96 lb. per sq. in. The average intake-pressure, as indicated 
by a gage in the suction pipe, was 4 lb. per sq. in. below atmospheric 
pressure. The pressure due to the hydrostatic head between the suction- 
and discharge-gages was 3.5 lb. per sq. in. What was the duty? 

Solution.— By For. (33),D A = 1,000,000 (P d ± Pi + P D ) AL f N s /H = 
1,000,000 X (96 + 4 + 3.5) X [20 2 X 0.7854 - (3 2 X 0.7854 -=- 2)] X 
(15 -r- 12) X 37,264 + 17,642,400 = 84,884,000 ft. lb. per 1,000,000 
B.t.u. 

50. The Miscellaneous Reciprocating-Pump Formulas 

which follow supplement those given previously herein. 
These formulas relate specifically to single-acting simplex 
pumps. The number of strokes per minute = the number of 
pumping strokes per minute = J£ the number of reversals of 
the piston. Where cylinder area is used in the following 
formulas, it means the cross-sectional area of the cylinder 
taken at right angles to the piston rod. 

Note. — In The Event That These Formulas Are Used In Double- 
Acting-Pump Computations, the number of working strokes per minute = 
the number of reversals per minute of the piston. Also, in double-acting- 
pump computations, for cylinder area must be substituted [cylinder area — 
(piston-rod area -5- 2)]. For (diameter of cylinder) 2 must be substituted 
{(diameter of cylinder) 2 — [(diameter of piston rod) 2 -=- 2} J. 

(34) Gal. per min. 

(Strokes per min.) X (Stroke in in.) (Di g, of water cyl. in in.) 2 

294 



Sec. 50] PUMP CALCULATIONS 35 

Example. — How many gallons of water will be delivered per minute 
by a pump having a water cylinder 8 in. in diameter by 12 in. stroke 
when it is making 100 strokes per minute? Solution. — Gallons per 
minute = (100 X 12 X 8 X 8) ■*- 294 = 261 gal per min. 

(35) Dia. of water cylinder in in. 



17. Uyj 



Gal. per min 



(Sirokeinin.) X (Strokes per min.) 

Example. — What will be the required water-cylinder diameter to 
pump 200 gal. per min., if the length of stroke is 10 in. and the pump 
makes 120 strokes per min.? Solution. — Diameter of water cylinder 
in inches = 17.14-\/(200) -*- (10 X 120) = 7 in. 

(36) Area of water cylinder in sq. in. 

(231) X (Gal per min.) 
(Strokes per min.) X Stroke in in.) 

Example. — What area of water cylinder is required to pump 330 gal. 
per min., if the pump has a 16 in. stroke and makes 80 strokes per min.? 
Solution.— Area of water cylinder = (231 X 330) -s- (80 X 16) = 59.6 
sq. in. 

(37) Area of water cylinder in sq. in. 

(3.85) X (Gal. per hr.) 



(Strokes per min.) X (Stroke in in.) 

Example. — A pump has a stroke of 24 in., and makes 50 strokes per 
min. What must be the water-cylinder area if it is to pump 97,920 gal. 
per hr.? Solution. — Area of water cylinder = (3.85 X 97,920) -s- 
(50 X 24) = 314 sq. in. 

(38) Length of stroke in in. 

(231) X (Gal, per min.) 

~~ (Strokes per min.) X (Area of water cyl. in sq. in.) 

Example. — What must be the length of stroke of a pump having a 
water-cylinder area of 28.3 sq. in., if it must pump 146 gals, per min. when 
making 120 strokes per minute? Solution. — Length of stroke = (231 X 
146) ^ (120 X 28.3) = 10 in. 

(39) Stroke in in. 

(Gal, per hr.) X (4.9) 

(Strokes per min.) X (Diam. of water cylinder in in.) 2 

Example. — What will be the required length of stroke to pump 35,251 
gal. per hr. if the pump has a water cylinder 12 in. in diameter and makes 
80 strokes per min.? Solution. — Length of stroke = (35.251 X 4.9) -f- 
(80 X 12 X 12) = 15 in. 



36 STEAM POWER PLANT AUXILIARIES [Div. 1 

(40) Stroke in in. 

(Gal, per min.) X (294) 

(Strokes per min.) X (Diam. of water cyl. in in.) 2 

Example. — What will be the required length of stroke to pump 587 gal. 
per min. if the pump has a water cylinder 12 in. in diameter and makes 
66.6 strokes per min. Solution. — Length of stroke = (587 X 294) -3- 
(66.6 X 12 X 12) = 18 in. 

(A\) Strokes T)er min — — 

(Water-cyl. area in sq. in.) X (Stroke in in.) 

Example. — How many strokes per minute will a pump have to make 
to pump 8,812 gal. per hr. if it has a water-cylinder area of 28.3 sq. in. 
and a length of stroke of 12 in.? Solution. — Number of strokes = 
(8,812 X 3.85) -T- (28.3 X 12) = 100 per min. 

(Gal. per hr.) X (4.9) 



(42) Strokes per min. = 



(Stroke in in.) X (Dia. of water cyl. in in.) 2 



Example. — How many strokes per minute will a pump have to make 
to pump 8,812 gal. per hr. if it has a water-cylinder diameter of 6 in. and 
a length of stroke of 12 in.? Solution. — Number of strokes = (8,812 X 
4.9) -s- (12 X 6 X 6) = 100 per min. 

(43) Strokes per min. 

_ (Gal, per min.) X (2 3 1) 

(Stroke in in.) X (Area of water cyl. in sq. in.) 

Example. — How many strokes must a pump make per minute to 
pump 146 gal. per min. if it has a water-cylinder area of 28.3 sq. in. and 
a 10 in. stroke? Solution. — Number of strokes = (146 X 231) -s- 
(10 X 28.3) = 120 per min. 

(44) Water-gage pressure necessary to balance steam-gage 

(Steam-qaqe pressure) (Diam. in in. of steam-cyl.) 2 

pressure = -~tjS =-^ — 3 : T^ 

^ (Diam. in in. of water-cyl.) 2 

Example. — If a pump has a steam cylinder 5 in. in diameter and a 
water cylinder 3 in. in diameter, what water-gage pressure will be re- 
quired to balance a steam-gage pressure 150 lbs. per sq. in.? Solution. 
— Water-gage pressure = (150 X 5 X 5) 4- (3 X 3) = 416 lbs. per sq. in. 

(45) Steam-gage pressure necessary to balance water-gage 

(Water-qaqe pressure) (Dia. in in. of water cyl.) 2 

pressure = - y y *. — r-^ — j—. ~ ^- L - 

(Dia. in in. of steam cyl.) 1 

Example. — If the water cylinder of a pump is 8 in. in diameter and 
the steam cylinder is 12 in. in diameter, what must be the steam-gage 
pressure in order to just balance a water-gage pressure of 130 lbs. per 
sq. in.? Solution. — Steam-gage pressure = (130 X 8 X 8) -f (12 X 12) 
<= 57.8 lbs. per sq. in. 



Sec. 50] PUMP CALCULATIONS 37 

(46) Area of water cylinder in sq. in. necessary to balance a 
given steam pressure = 

(Area of steam cyl. in sq. in.) X (steam pressure in lbs, per sq. in.) 
(Water pressure in lbs. per sq. in.) 

Example. — A pump has a steam-cylinder area of 113.1 sq. in. If the 
steam gage reads 60 lbs. per sq. in. and the water-pressure gage reads 
135 lbs. per sq. in., what must be the area of the water cylinder if the 
piston is just balanced? Solution. — Area of water cylinder = (113.1 X 
60) + 135 = 50.25 sq. in. 

(47) Area of steam cylinder in sq. in. necessary to balance a 
given water pressure = 

(Area of water cyl. in sq. in.) X (Water pressure in lbs. per sq. in.) 
(Steam pressure in lbs. per sq. in.) 

Example. — A pump has a water-cylinder area of 50.25 sq. in. If the 
water gage shows a pressure of 135 lbs. per sq. in. and the steam gage 
shows a pressure of 60 lbs. per sq. in., what must be the area of the 
steam cylinder if the piston is just balanced? Solution. — Area oj 
steam cylinder = (50.25 X 135) 4- (60) = 113.1 sq. in. 

QUESTIONS ON DIVISION 1 

1. What conditions govern the height to which water may be lifted by pump-suction? 
What is the practical limit of suction-lift at sea-level? What is the practical limit of 
temperature at which water may be lifted by pump-suction? 

2. What is a static head? What is its significance? 

3. Why should water from an open heater enter the suction-nozzle of a boiler feed 
pump under a static head? Describe the action that may occur within the pump 
cylinder, if the inlet static-head is insufficient. 

4. Enumerate the three general forms of resistance, or head, which must be overcome 
in pump-operation. Which of these comprise the dynamic head? 

5. What is velocity-head? Friction-head? Measured-head? 

6. Enumerate the causes of friction-head. 

7. What is entrance-head? 

8. If a pump is discharging into the compression-tank of an elevator system, how is 
the head due to the gage-pressure in the tank classified in computations relating to the 
performance of the pump? 

9. What is the total head on a pump? 

10. Do computations based upon values taken from published tables afford, in all 
cases, accurate criteria of the water-friction in pipes? Why? 

11. What is the displacement of a reciprocating pump? 

12. What constitutes the effective displacement area of an outside-end-packed 
plunger? Of a center-packed plunger? Of an inside-packed plunger or piston? 

13. What is pump-slip? Under what circumstances may pump-slip be negative? 

14. Explain the influence of high piston speed on pump-slip. 

15. What is meant by the volumetric efficiency of a pump? 

16. What should be the maximum limit of piston-speed for a pump with a 20-in. 
stroke? With a 9-in. stroke? With a 3-in. stroke? 

17. What is meant by the useful work of a pump? The actual work? 

18. What is meant by the indicated efficiency of a reciprocating pump? What losses 
does this efficiency particularly signify? 



38 STEAM POWER PLANT AUXILIARIES [Div. 1 

19. What is meant by the hydraulic efficiency cf a pump? 

20. What constitutes the total head in determining the hydraulic efficiency? 

21. Describe an experimental method of determining the load on a pump. 

22. What is meant by the mechanical efficiency of a reciprocating pump? What loss 
is determined by this efficiency? 

23. What is meant by the total efficiency of a pump? What does this efficiency 
signify? 

24. What is meant by the duty of a steam pump? 

25. What conditions may vitiate the practical significance of pump-duty computed 
on a basis of coal-consumption? On a basis of steam-consumption? 

26. Wherein lies the practical value of pump-duty computed on a basis of heat- 
consumption? 

PROBLEMS ON DIVISION 1 

1. Atmospheric pressure at an altitude of 13,000 ft. above sea-level is approximately 
9 lb. per sq. in. What is the practical suction lift at this elevation? 

2. A direct acting steam pump is lifting water through a height of 11 ft. and dis- 
charging it through an additional height of 19 ft. What is the total static head, ex- 
pressed in terms of pressure? 

3. A boiler feed pump has the water fed to it (Fig. 3) by gravity. It is assumed that 
the inlet head thus produced is wholly expended in filling the pump cylinder with water 
against a tendency of the water, due to its temperature, to vaporize in the cylinder. 
Hence no part of this head is available for balancing an equivalent head on the delivery 
side. The delivery pipe is of 1J£ in. size. It has a total horizontal length of 115 ft. 
and a vertical length of 38 ft. There are three 90 deg. elbows, two plugged tees and 
two globe valves in the fine. The boiler pressure is 150 lb. per sq. in. If about 20 gal. 
of water are delivered per minute, what pressure head will be necessary in the pump 
cylinder? What is the equivalent gage pressure? 

4. If all conditions remain the same as in prob. 3 except that the pipe-size is changed 
to 1 in., how many gallons of water will be delivered? 

5. A direct-acting simplex steam pump is required to deliver 90 cu. ft. of water per 
min. The flow velocity in the suction pipe is assumed to be 210 ft. per min. and in the 
discharge pipe 390 ft. per min. What should be the sizes of the piping for suction and 
discharge? 

6. An outside end-packed duplex plunger pump has plungers of 10 in. diameter. 
The stroke is 20 in. Each plunger makes 65 strokes per min. What, if the pump is 
double acting, is the displacement in cubic feet per minute? 

7. The displacement of a pump is 510 cu. ft. per min. The pump delivers 487 cu. 
ft. of water per min. What is the slip? 

8. What is the volumetric efficiency of the pump of Prob. 7? 

9. The plunger diameter in a direct acting simplex steam pump is 3.5 in. The stroke 
is 6.5 in. When the plunger makes 110 strokes per minute, the volumetric efficiency is 
98 per cent. What is the discharge? 

10. A direct acting duplex steam pump is required to deliver 990 cu. ft. of water per 
hr. while running 100 ft. of piston travel per min. If the volumetric efficiency is 96 
per cent, what should be the water-piston diameter? 

11. The water piston diameter in a direct-acting steam pump is 5 in. The pump 
discharges through a 2 in. pipe. The piston travel is 80 ft. per min. What is the 
velocity of flow in the discharge pipe? 

12. A pump elevates 20,106 lb. of water per minute through a total vertical height 
of 38.5 ft. What net work, in foot pounds, is done in one minute? 

13. What is the net horse power expended by the pump in Prob. 12? 

14. What is the horse power required for lifting 9,500 lb. of water per minute against 
a useful head of 310 ft. when the total efficiency of the pump is 85 per cent.? 

15. A steam pump elevates 9,000,000 lb. of water against a total useful head of 120 ft. 
The coal consumption of the boilers while furnishing steam for this work is 3,500 lb. 
What is the duty of the pump per 100 lb. of coal? 



DIVISION 2 
DIRECT-ACTING STEAM PUMPS 



51. Direct-Acting Steam Pumps For Modern Power-Plant 
Service are (Fig. 27) of the reciprocating double-acting, 
suction type. That is, they are designed to raise water by 
suction from a lower level, and to deliver it during each stroke 
of the moving element (Fig. 28) to tanks, boilers, or wherever 
else required. 



Metal Snap- 
Rings-^ 




Brass Liner Forced • 

into Cylinder Under } 

Pressure 



Water-Piston 



^\\\v\\\\^\\\v\\m\\\\\m 



Fig. 



27. — Water-End Of Direct-Acting Steam-Pump Having Water-Piston Fitted With 
Snap Rings. 



Explanation. — The movement toward the left of the piston (P-Fig. 
28) as indicated, causes the water in the space B to be forced out through 
the left-hand pair of discharge valves, Vd. Coincidentally, it creates a 
partial vacuum in the space A. That is, it causes the air pressure in 
the space A to be lowered. This reduction of pressure, per square inch, 
must be equal to, or greater than, the pressure per square inch which is 
imposed by the weight of a column of water of the height Lhs of Fig. 1. 
The external atmospheric pressure will then force the water up the suc- 
tion pipe, S, and through the right-hand pair of suction valves, V s . On 
the return stroke, a partial vacuum is created in the space B. Water 
then enters space B through the left-hand pair of suction valves, while 
the water in A is forced out through the right-hand pair of discharge 
valves. 

39 



40 



STEAM POWER PLANT AUXILIARIES 



[Div. 2 



Note. — The intake water often flows under pressure to the suction 
nozzles (Sec. 4) of power-plant pumps. The intake pressure may be 
due (Fig. 29) to an elevated source of supply, or it may be derived from 
street mains. 



Discharge Pipe 

cylinder-. x^\\\\\^ l 




Suction Pipe- ......... .^Jnb-I Suction Valve 

Fig. 28. — Illustrating The Principle Of 
The Reciprocating Pump. 



■ ' N ■•"=■• ■ -- suctwri from Well: 1 

Fig. 29.— Intake Water Often Flows Un- 
der Pressure To The Suction Nozzle. 



52. The Allowable Velocity Of Flow In The Water-Piping 
Of A Direct-Acting Steam -Pump is: (1) For the intake-pipe, 

Discharge- 
Packing... : Nozzle— + 




Suction Nozzle '■ 
Fig. 30. — Outside Center-Packed Plunger-Pump. 

about 200 ft. per min. (2) For the discharge-pipe of a single 
pump, about 400 ft. per min. (3) For the discharge-pipe of a 
duplex pump, about 500 ft. per min. (4) For the centrifugal 



Sec. 53] 



DIRECT-ACTING STEAM PUMPS 



41 



pump about 600 ft. per min. in both the discharge and suction 
pipes. 



.-Plunger 



■Cylinder 




Fig. 31. — An Outside End-Packed Plunger-Pump. 

53. Direct-Acting Steam Pumps May Be Classified, With 
Reference To Their Water-Ends, as follows: (1) Piston- 
Pumps (Fig. 27). (2) Plunger-Pumps (Fig. 30). The latter 



Eye-Bolt- 



Discharge 
Outlet-, 



Brass Liner 
Secured 

with •-. 
Cap-Screws '< 



Discharge ^tSSSxl -fc^S 

Nozz.'e--- 




Suction Nozzle- 



ww^^s^^^m^- 




Fig. 



32. — Pump-Plunger Inside-Packed 
With Metal Ring. 



•Suction Inlet 



Fig. 33.— Water-End Of Direct-Act- 
ing Steam-Pump With Fibrous-Packed 
Water-Piston. 



may be subdivided into: (a) Outside end-packed plunger-pumps 
(Fig. 31). (b) Outside center-packed plunger-pumps (Fig. 30). 
(c) Inside-packed plunger-pumps (Fig. 32). In a piston-pump, 
the piston traverses a liner or barrel (Fig. 33) which is com- 



42 



STEAM POWER PLANT AUXILIARIES 



[Div. 2 



monly made of brass. The liner may be secured (Fig. 27) by 
means of a force-fit with the bore of the iron cylinder casting, 
or (Fig. 33) by means of stud-bolts or cap-screws. A tight 
joint between the periphery of the piston and the bore of the 
liner is obtained (Fig. 33) by means of rings of square fibrous 
packing, or (Fig. 27) by using metal snap-rings. In a plunger- 
pump either end-packed, center-packed or inside-packed, the 
plungers pass through fibrous — or metal-packed stuffing-boxes. 

Note. — Piston-Pumps May Be Used Against Discharge-Heads 
Up To About 300 Lb. Per Sq. In. (Sec. 38). Difficulty may be had, 
however, in keeping the piston tightly packed if the head-pressure ex- 
ceeds 150 lb. per sq. in. The fact that the packing is stationary in the 



Pump Cylinder : Pump Cylinder 
Liner-. '« Casting? 




Pump Cylinder .-Pump Cylinder 
Liner-, \ Castinq 




■Binding Nut ' -Piston Pmg/s 
Fig. 34. — Metal-Packed Pump-Piston. 



'-Binding Nut 'Water Grooves 
Fig. 35. — Water-Packed Pump-Piston. 



plunger pumps and that it may be tightened more readily, renders it 
more effective therein than in piston pumps. For low pressures, the 
piston are less expensive than the plunger pumps. 

Plunger-Pumps Are Commonly Used Against Discharge-Heads 
Above About 200 Lb. Per Sq. In. For pressures above 300 lb. per sq. 
in., choice of plunger-pumps as against piston pumps, is practically 
imperative. 

Water-Pistons Packed With Metallic Rings (Fig. 34) are com- 
monly used in hot-water pumps. Water-packed pistons (Fig. 35) are 
also sometimes used for hot-water service. The packing is afforded by 
the water which becomes pocketed in a series of annular grooves in the 
piston's periphery. 

54. The Water-Piston Packing In Direct-Acting Steam 
Pumps For Power-Plant Service is generally fibrous. It is 
commonly known as canvas or duck hydraulic-packing. It 
consists mainly of cotton fiber (Fig. 36) interlaid with a 
rubber composition. Its cross-section is square. 



Sec. 55] 



DIRECT-ACTING STEAM PUMPS 



43 



Note. — Rings of Canvas Packing For a Pump-Piston should 
(Fig. 36) be cut about ${$ in. short of meeting when inserted (Fig. 37) 
in the cylinder-bore. Also, the joints should be lapped (Fig. 36). This 
packing is commonly made in layers. The layers can (Fig. 38) be peeled 
off to get rings of suitable width or thickness. If the packing is too deep 



Top View of 
■fucking Riny 




Flcmo/ed End Sronze cufmder ToKtina P/ '^ e - 
of Piston l Limr y' c y" mercc * STin 9 Stick 




Packing Recess 



Fig. 36. — Ring Of Tuck Canvas Fig. 37. — How Packing Is Inserted In Packing 
Piston-Packing. Recess Of Pump-Piston. 

to fit the recess around the piston, it may be cut down. A convenient 
and accurate method (Fig. 39) of doing this is by gripping the packing 
in a vise and paring it with a sharp drawing knife. The rings should be 
well coated with graphite and cylinder oil. They should be just tight 
enough to require moderate pressure of the fingers to force them into 
the recess around the piston. They may then be forced home (Fig. 



Cylinder, : Rkicf Peeled Down 
Cas ting* V. to Fit 




RmascfTuckis -Follower 
cklncf....' Plate 

Fig. 38. — A Canvas-Packed Pump 
Piston. 



Canvas Packina- 


y/fl 


i^^^g 


1181^5^31^ 




Jill 


Vise-- ' JA 


f 1 Drawing'' 
Knife 



Fig. 39. — How Depth Of Canvas Piston-Packing 
May Be Cut Down. 



37) with a stick of soft wood. Rings of canvas packing may be partially 
expanded to their working size, before inserting them in the piston-recess, 
by soaking them for a few hours in warm water. 



55. The Valves Of Power -Plant Pumps are generally of the 
poppet-disc type (Figs. 40, 41, 42, and 43), rising vertically 
from flat seats. Conical-seated valves (Fig. 44) are also used. 



44 



STEAM POWER PLANT AUXILIARIES 



[Div. 2 



The discs (D-Fig. 40) of flat-seated valves are commonly made 
of a composition of rubber with certain other substances. 
They are also made of metal (Fig. 43). Usually a brass cap 



Lock-Nut- — -p^ 


gi^L*; . -Spring-Nut 


Brass Cap-, (~~~j 


jpjfpwp /Spring 


.■-Composition \ T2E 
Rubber Disc V_)^ 


3Ml£>^J. Cast Iron- 


Y--'-^^Il21IH| 


'; ffi^T^3\ ^\ 


fc ' '-'^o/; V s 4l 


Il§c7 ^Kllll^M 


i SMI 


=^^/ ~\Xc^ ^ 


Brass Seat-Bushing^ 


^'"Brass Valve-Stud 



Fig. 40.— Flat-Seated Pump- Valve With 
Composition Rubber Disc. 



Stud--' 




Fig. 41.— Rubber Pump- Valve Flat-Disk 
Type. 



Brass- 
Valve 
Disc 




Fig. 42. 



-Bronze Pump- Valve Flat- 
Disk Type. 



Stud- 
Spring- ... 
Valve Disc-, 




Fig. 



43. — Sectional Elevation Of Bronze 
Pump-Valve. 



.-Spr ing Follo wer ..-Lock - Nut 
Spring 



Conical-. 
Sew/- • 




. Brass 
Brass Seat \ Disc 
Bushing--' 



Brass Valve 
Stuol 



Cast-Iron, 

Valve Deck 



Fig. 44. — Conical-Seated Pump-Valve With Brass Disc. 

or plate (P-Fig. 40) is used to stiffen the rubber disc and 
prevent warping. It also serves to protect the disc from 
the direct thrust of the spring (£-Fig. 40). The discs of 
conical-seated valves (Fig. 44) are generally made of metal. 



Sec. 55] 



DIRECT-ACTING STEAM PUMPS 



45 



Note. — The Hardness Of Rubber Composition Valve-Discs should 
be adapted to the special requirements of the service for which the discs 
are intended. The valve-discs of vacuum pumps should be soft and 
pliable. Such discs are also suitable for pumps working against water 
pressures up to about 75 lb. per sq. in. For pressures from about 75 
to 150 lb. per sq. in., hard rubber composition discs usually give the best 



Wing- Type Discharge Valves-. . ^^^e v --' ^x. 



•■Outlet 



Diaphragm 




\\\\\ \\\\\ \ > www w\. w \ \ \ w \ \ www 

Fig. 45. — Water-End Of Direct-Acting Steam-Pump For Hydraulic-Pressure Service* 



service. For pressures from about 150 to 300 lb. per sq. in., specially- 
hard vulcanized rubber composition valve-discs generally suffice. Metal 
valve-discs are required for pressures above about 300 lb. per sq. in. 
The hardness of valve-discs should also depend on the temperature of 
the water pumped. The higher the temperature of the pumped water, 



Left-hand Side • Nozzj^^ Right Hand-Side 




A\\w\ 



Packing Gland- 



Fig. 46. — Water-End Of Duplex Outside- 
Packed Plunger-Pump Equipped With 
Pot Valves. 




Seat Bushing- 



Fig. 47.— Ball Pump-Valve For High 
Pressure Service. 



the harder the valve discs should be. Metal valve discs are frequently 
used for hot-water service. 

The Seats Of Metal-Disc Pump-Valves Should Be Of The 
Same Kind Of Metal As The Discs. This is to prevent electrolytic 
action. 

Wing-valves are commonly used in high-pressure pumps (Figs. 45 and 



46 



STEAM POWER PLANT AUXILIARIES 



[Div. 2 



46) of the pot-valve type. Ball-valves (Fig. 47) are also sometimes used 
in pumps of this type. A valve which is commonly used in pumping 
clear liquids is shown in Fig. 48. 



Composition, 
.■Valve Disc 



■Stem 

Conical Spring 




Fig. 



48. — Type Of Pump Valve 
Used For Clear Liquids. 



Fig. 49. — Kinghorn Pump Valve. 



56. The Kinghorn Valve For Air Pump Service (Fig. 49) 
consists of three bronze disks, each about 3^32 in. thick, which 
are mounted loosely on a central stud. Buckling and distor- 
tion of the discs is prevented by a guard which limits the lift. 

57. Pump-Valve Seats May Be Either Forced Or Threaded 
Into The Valve Decks (Figs. 50 and 43). Where the seat is 



Pubber 



..'Spring 



I. Va/re 




Taper J/t- 



Ya/re Seaf--' 71 
Peaned Edge''' 

Fig. 50. — Rubber Valve For Low Pressure 
Warm- Or Cold-Water Service. 




Fig. 51. — Flat-Faced Wing Poppet Valve. 



forced into the deck, the hole in the deck is bored to a very 
slight taper, and the cylindrical portion of the seat is turned to 
correspond. When the seat has been forced in, the projecting 
edge, E (Fig. 50) is peaned over to prevent the seat from work- 



Sec. 5$] 



DIRECT-ACTING STEAM PUMPS 



47 



Discharge- 
Valves 



Suet ion 
Valves 



ing out. Where the seat is threaded into the valve deck, the 
threads are turned on a slight taper to insure a tight fit. 

58. Flat-Faced. Bronze Poppet Valves (Fig. 51) are used 
on pumps of the pot-valve type (Fig. 46) for high pressures. 
The vertical movement of these valves is guided by wings 
which work in the valve-seat openings. 

59. Three Different Methods Of Arranging The Valves 
Of Horizontal Double-Acting Suction Pumps are in use: 
(1) The sets of discharge- and of suction-valves may be super- 
imposed one above the other (Fig. 52) above the pump-barrel 
or cylinder. (2) The sets of suction- and discharge-valves may 
be arranged (Fig. 53) side 
by side above the pump- 
barrel or cylinder. (3) 
The discharge-valves may 
be located (Fig. 1) above 
the pump-barrel or cylin- 
der and the suction-valves 
below. 

Arrangement (1) is 
commonly used in small 
low- or medium-pressure 
pumps. It admits of 
easy access to the valves 
for renewals and repairs. 
Its disadvantage is that 
it (Fig. 52) requires a 
reversal of the flow of 
water through the pump. 
This tends to a dimin- 
ished pumping capacity. A pump having this arrangement 
is termed a submerged-piston pump. Practically all small 
boiler-feed pumps (Sec. 198) and wet vacuum pumps (Sec. 
1353) are so constructed. 

Arrangement (2) is commonly used in pumps (Fig. 53) 
designed for high pressures. It permits a structural design 
which is conducive to great strength. 

Arrangement (3) is much used in large low- or medium- 




( ■> \ } Inlet-' 



Fig. 52. — Medium-Pressure Piston-Pump With 
Suction And Discharge Valves Arranged Above 
Pump-Barrel. 



48 



STEAM POWER PLANT AUXILIARIES 



[Div. 2 



pressure pumps. It permits the water to pass through the 
pump without any reversal of flow. 



Suction Valve- 



j. ■ • -Discharge Orifice 
-Valve Pots 



Suction 
-Orifice 



Pump 
.■■Barrel 




Cross head--' 



mV^^\\\\\\^ 



Fig. 53. — Outside-Packed-Plunger High-Pressure Pump With Suction- And Discharge- 
Valves Arranged Side By Side Above Pump-Barrel. 

60. The Total Effective Area Of Opening Of Each Set Of 
Suction- And Discharge -Values In A Direct-Acting Steam- 




Discharge 
Valve-. 



.-Discharge 
I Valve 
' Chamber 




Discharge 
Valve 
Deck 



Fig. 54. 



-Open Position Of Flat-Disk 
Pump-Valve. 



Fig. 55.— Water-End Of Single Direct- 
Acting Steam-Pump With Air-Chamber. 



Pump should, for low speeds, be about 30 per cent., and for 
high speeds, about 50 per cent., of the piston- or plunger area. 



Sec. 61] 



DIRECT-ACTING STEAM PUMPS 



49 



Note. — The Area Of Opening Given By A Flat Disc Valve (Fig. 
54) is the annular area obtained by: Multiplying the lift, L, in inches, by 
the diameter, d, in inches, and by 3.14. The most adaptable valve- 
diameter has been found to be from 3 to 4 in. The lift commonly used 
is about ¥± in., regardless of the water-diameter. 

Example. — The water-piston of a high-speed pump is of 10-in. diame- 
ter. The piston-rod is of 3-in. diameter. How many flat disk valves, 




\\\\\\\\\\\\\\\\\\\\\\\\\\W: V . 

Compressor Discharge-Pipe-- - 
Fig. 56. — Apparatus For Replenishing Air-Chamber In Discharge-Pipe Of Hydraulic 
Elevator Pump Under 800 Lb. Pressure Per Sq. In. 

each of 3-in. diameter and 3^-in. lift, are required for each set of suction- 
and delivery-valves in this pump? 

3 2 X 7854 
Solution. — The effective piston-area = (10 2 X 0.7854) k~ 

= 75 sq. in. The area of opening of each valve will (Sec. 60) be 0,25 X 
3 X 3.14 = 2.36 sq. in. Hence, (Sec. 60) 75 X 0.5 ^ 2.36 = 15.9, or, 
practically, 16 valves are required in each set. 

61. Air-Chambers are often connected to the discharge- 
valve chambers (Fig. 55), or to the discharge pipes (Fig. 56), of 



50 



STEAM POWER PLANT AUXILIARIES 



[Div. 2 



Check Valves 
Both Opening To- 
ware/ Air- Chamber- 



Chamber- 



direct-acting steam-pumps. The function of an air-chamber is 
to provide a cushion for the discharged water. 

Explanation. — The air in the chamber, C, (Fig. 55), is compressed, 
during discharge, to a pressure approximately equal to the pressure 
against which the pump is working. Thus, it forms a highly elastic 
buffer or cushion. When the piston reaches the end of its stroke, the 
discharge suddenly ceases. An instant elapses before the opposite 
stroke begins. During this instant, expansion of the compressed-air 
in C tends to keep the discharged water in motion. Hence, the reacting- 
tendency of the column of water in the discharge pipe is neutralized. 
Consequently, where the air-chamber is of the proper proportions, no 
shock, neither to the piping nor to the pumping mechanism should result 
therefrom. 

Note. — The Air-Chambeb Is a 
Less Needful Accessory in Du- 
plex-Pump Service Than In Sim- 
plex-Pump Service. Duplex-pumps 
(Sec. 71) have continuous piston- 
travels. Hence, with such pumps, the 
discharge of water is approximately 
continuous. For high -pressure 
duplex-pumps (Sec. 75) and those 
working against very high pressures, 
as in high-pressure hydraulic elevator 
service, air-chambers are, neverthe- 
less, distinctly necessary. 

62. The Height Of The Water- 
Level In The Air-Chamber Of A 
Pump (Fig. 55) should not ex- 
ceed one-fourth the height of the 
chamber. With small slow- 
running pumps, working against 
pressures below 50 lb. per sq. 
snifter For Replenishing in., it is usual to rely entirely 

Air-Chamber Of Direct-Acting Steam- upon t ^e a i r _bubbleS, which are 
Pump. . . ' 

entrained with the suction- 
water, for maintaining the requisite volume of air in the 
chamber. Where pumps run at high-speeds and against 
pressures higher than about 50 lb. per sq. in., good service 
requires that the air-chambers be recharged occasionally by 
mechanical means, or by use of a snifter. The snifter (Fig. 
57) may be operated by the pump itself. It is suitable for 




Dischctrgfi 

Stroke 

Suction Stroke 

Plane of Suction-Valve Deck 



Fig. 57. 



Sec. 63] DIRECT-ACTING STEAM PUMPS 51 

pumps running against pressures up to about 200 lb. per sq. in. 
It can be used only where the pump has a suction-lift. 

Explanation. — The snifter is connected to the pump-cylinder at a 
point, P, (Fig. 57) between the suction- and discharge valve-decks. 
When the valves V and Vi are opened, water is forced, during the 
head-end discharge-stroke of the pump-piston, into the snifter-cylinder, 
S. The air in S is thus dislodged and forced into the air-chamber, A, 
through the check-valve C. During the corresponding suction-stroke, 
the water in S is drawn back into the pump cylinder. Thus the snifter- 
cylinder is again filled with air through the check-valve C\. 

The flow through valve V\ should be throttled on the suction-stroke 
to prevent all of the water from being drawn from cylinder S. The 
purpose of this is to retain a column of water in S to act as a piston for 
driving the air through check-valve C. Valve Vi should be so ma- 
nipulated as to establish a regular pulsation, within the length of the 
glass gage, of the water-level in S. 

63. Air-Chamber Charging-Apparatus For Pumps Working 

Against Very High Pressures, usually depend (Fig. 56) for 
their effectiveness, upon the tendency of particles of com- 
pressed- air to percolate through masses of water. 

Explanation. — Gate valve V (Fig. 56) being closed and V2 opened, 
the air compressor, C, is started. Gate valve V\ is then opened to 
permit the water in the reservoir, R, to be blown out, after which it is 
closed. When the pressure within the reservoir reaches the limit of the 
compressor's capacity for compression, which may be about 75 lb. per 
sq. in., valve V2 is closed and V is opened. Water then passes through 
the connecting-pipe and gradually fills reservoir R. Coincidentally, the 
compressed air, thus displaced, bubbles through the water in the con- 
necting-pipe and upward through the mass of water in the lower part 
of the air-chamber, A. The gage-cocks, G, are used to determine the 
approximate height of the water in the air-chamber, A. 

64. The Ratio Of Air-Chamber Volume To Volume Of 
Water -Piston Displacement In Direct-Acting Steam Pumps 

may, for ordinary rates of speed, be about as follows: (1) For 
simple pumps (Fig. 57) from 2 to 3.5. (2) For duplex pumps 
(Fig. 58) from 1 to 2.5. The air-chamber volume of a pump 
for high-speed service (Fig. 25) may be from 5 to 6 times the 
volume of piston displacement. 

65. Vacuum Chambers (Figs. 59 and 60) are sometimes 
attached to the suction-pipes of direct-acting steam-pumps. 
The function of a vacuum-chamber is to insure that the pump- 



52 



STEAM POWER PLANT AUXILIARIES 



[Div. 2 



Air 

Chamber 




y///////s////ys/s^ 



Fig. 58. — Showing Height Of Water 
In Vacuum-Chamber At Instant Of 
Piston-Reversal. 



Air S~\ 

Chamber- • - 



^Discharge 
.■■Outlet 



Suction 
Inlet to 
Pump 



Suction 
Pipe-., 




Fig. 59. — Vacuum-Chamber Con- 
nected To End Of Suction Pipe Of 
Direct-Acting Steam-Pump. 



Duplex 
Steam 
Pump-*, 




OV Vacuum -, 
<§>] Chamber 



^W^^WW^ 



Wafer-Level at 

Reversal of 

Stroke- - 



Water- Level 

During — - 

Travel of 

Both Pistons 

Suction-Pipe, ' 

Fig. 60. — Special Form Of Vacuum-Chamber. 



Sec. 66] 



DIRECT-ACTING STEAM PUMPS 



53 



cylinder be completely filled with water at each reversal of 
the piston-stroke. It also provides an air-cushion for the 
column of water in the suction-pipe when the movement of 
the water is suddenly arrested, due to the momentary stoppage 
of the piston at the end of each stroke. 

Explanation. — During the piston-stroke the air (Fig. 58) in the 
vacuum-chamber tends (Fig. 61) to expand. Therefore, if the current of 
water in the suction-pipe is insufficient to completely fill the space behind 
the piston, a portion of the water standing above the plane, X Y, of the 
suction-inlet is forced into the cylinder. Thus, the cylinder will be full 
of water when the piston-stroke is reversed. When the flow of water 
(through the suction valves) momentarily ceases at the end of the stroke, 



Vacuum 
Chamber 



Air Chamber 




Fig. 61. — Showing Height Of Water In Vacuum-Chamber During Progress Of Piston 

Stroke. 



the momentum of the moving column in the suction-pipe is expended in 
compressing (Fig. 58) the air in the vacuum chamber. Thus the shock 
that might otherwise attend abrupt stoppage of the flow is avoided. 

66. Direct-Acting Steam-Pumps May Be Classified, With 
Reference To Their Cylinders, as follows: (1) Single or 
simplex pumps. (2) Duplex pumps. A simplex pump (Fig. 
62) has one steam-cylinder and one water-cylinder. A 
duplex pump (Figs. 63 and 64) has two steam cylinders and 
two water cylinders. It comprises, in effect, two single 
pumps, A and B, (Fig. 63) placed side by side, drawing water 
through a common suction-pipe, S, and discharging into a 
common delivering chamber, C, and pipe D. 



54 



STEAM POWER PLANT AUXILIARIES 



[Div. 2 




^^il^^^^fcs; 



Fig. 62. — Longitudinal Sectional Elevation Of Burnham Direct-Acting Simplex Steam- 
Pump. 



-Left-Hanoi Side ,.-Abutments 



.-■—Piston-Rods 






jj MgfarTS 




\ Delivery ■§ 

•Right-'- \ I'Plan View Of Valve Gear, Slide-Valves Central \-Chamber gj 

Z/be 




mmw\\wmw\\\m^^ 



U. • Elevation Of Valve Gear, Rocker- Arms Perpendicular 
Fia. 03. — Plan And Elevation Of Valve Gear Of Duplex Steam Pump. 



Sec. 6^ 



DIRECT-ACTING STEAM PUMPS 



55 



Note. — Each Steam-Valve Of a Duplex Pump Is Actuated By 
The Opposite Piston-Rod. The reciprocative motion of the piston- 
rods, Ri and R 2 , (Fig. 63) is transmitted to the slide-valves, Vi and V 2f 
through a system (Fig. 65) of oscillating rocker-shafts and arms. 




67. The Steam-Valve Gears Of 
Simplex-Pumps are (Figs. 62 and 
66) variously constructed. With 
all forms of such gears, however, 
the main valve for admitting steam 
to the cylinder and releasing it 
therefrom, is operated by direct 
steam-pressure. The valve is thus 
said to be steam-thrown. 



Rocker- Sfantf 




Fig. 64. — Sectional Elevation Of 
One Side Of Vertical Duplex Steam 
Pump For Boiler Feed Service. 



-Phton-Rools-- 



Fig. 65. — End-View Of Steam-Valve-Actuat- 
ing-Mechanism Of Duplex Pump. 



Explanation. — At the beginning of the inboard stroke (Fig. 62), 
main steam-port E is covered by the piston, P. Enough steam to give 
the piston an easy start passes in behind it through pre-admission port 
d (Figs. 62 and 67). When the piston moves far enough to uncover 
port E, it receives, through the valve-port V 2 (Figs. 62 and 67) and main 
steam-port E, the full steam pressure. It then moves at normal speed 
until it covers main steam port F (Fig. 62). By this covering of port F } 
the exhaust steam ahead of the piston is trapped in the inboard end of 



56 



STEAM POWER PLANT AUXILIARIES 



[Div. 2 



Lubricator Connec 



Steam 
Chest 
.Crank- 
Handle 



the cylinder. The exhaust steam thus forms a cushion, against which the 
piston makes an easy stop. 

During the inboard stroke, the actuating lever, A, is shifted to the 
opposite angular position, as indicated (Fig. 62) by the dotted lines. 

The toe of the actuating lever 
thus strikes tappet-block K and 
shifts the auxiliary slide-valve, 
H (Fig. 67) far enough to the 
left to open communication 
between auxiliary steam-port 
C 2 and auxiliary exhaust-port 
R. Coincidentally, the auxiliary 
valve, H, will admit live-steam, 
through auxiliary port d, to 
the right-hand end of chest- 
piston, M (Figs. 62 and 67). 
This will cause the chest-piston, 
which engages with and shifts 
the main slide-valve D, to move 
instantly to the left. Thus the 
main steam-port F will be 
brought to coincide with the 
drilled ports, V h in the main 
slide-valve. Preadmission port 
(r 2 will likewise be open. Steam will thus be admitted to the right- 
hand end of the cylinder for a reversal of the piston-stroke. 

If steam is admitted to the steam chest, M, (Fig. 66), it will enter the 
hollow ends, H, of the steam-chest plunger, F, and issue through a hole 

-----:r--T'' 5f !T^ X!l! ^r° rf %?y'^'- Actuatlng-lever Pivot-Stud, 
\ ^Auxiliary ExhausWort^^^wuhonry Slide Valve } * fTTt"" ' ' 




Reversing Valve. 



Exhaust Cavity 



wmsmtmmmksmm 



Fig. 66. — Sectional Elevation Of Steam-End 
Of Cameron Direct-Acting Simplex Steam- 
Pump, Showing Inside-Operated Valve 
Mechanism. 




Fig. 67. — Plan Of Steam-Valve Gear Of Burnham Direct-Acting Simplex Steam-Pump. 



in each end. The spaces between the ends of the plunger, F, and the 
heads of the steam chest will thus be filled with steam. Steam will also 
enter the cylinder through the port Pi and drive the main piston, C, 
to the left. When the main piston strikes the stem of the reversing 



Sec. 68] DIRECT-ACTING STEAM PUMPS 57 

valve R2 and forces this valve to the left, the steam at the left-hand end 
of the plunger, F, will escape through the port, E 2 , into the annular 
cavity A 2 and thence through a cored passage (not shown) in the cylinder 
casting into the exhaust cavity, K. The balance of pressure between the 
two ends of the steam-chest plunger, F, will thus be destroyed. Due to 
the preponderence of pressure at the right-hand end, the plunger will be 
instantly thrust to the left-hand end of the steam chest. The slide valve, 
D, is attached to the plunger, F. Hence it will likewise be shifted to the 
left. Live steam will then enter the left-hand end of the cylinder through 
the port P 2 , while the spent steam in the right-hand end will be exhausted 
through the port Pi. 

Instantly, when the main piston, C, starts on the return stroke, the 
reversing valve, R 2 , will be closed by the pressure of the steam which is 
constantly in contact with it through the dotted port S 2 . When the 
main piston has traveled far enough to the right, it will shift the reversing 
valve Ri. The series of events described above will then be repeated 
at the right-hand end. 

68. The Length Of Stroke Of A Simplex Pump Having 
External Value Gear (Fig. 62) depends upon the adjustment 
of the auxiliary slide-valve, H, (Fig. 67). 

Explanation. — Prick-punched shop-marks on the tie-rod, X, which 
forms the bearing for the piston-rod guide (Fig. 62) indicate the extreme 
travel of the piston in each direction. If the inboard stroke is too short 
it may be lengthened by a slight shifting of the tappet-block L (Fig. 
62), along the valve-stem, toward the right. The outboard stroke may 
likewise be lengthened by shifting the tappet-block K toward the left. 
These adjustments will permit the actuating-lever, A, (Fig. 62) to oscil- 
late further in each direction before striking the tappet-blocks. Shifting 
of the auxiliary valve H, (Fig. 67) will thus be delayed. 

If the piston-rod guide travels very close to the marks, the piston may 
hesitate before reversal at the end of each stroke. Or, it may sometimes 
hang at the end of a stroke. When this occurs the tappet-blocks, K 
and L, (Fig. 62) should be shifted closer to the toe of the actuating lever. 

69. Adjustment Of The Steam-Valve Of A Direct-Acting 
Duplex -Pump (Fig. 63) consists, first, in plumbing the rocker 
arms and setting both valves line-and-line with the outer 
edges of the steam-ports. The lost-motion, or clearance, 
between the valve-stem collars and the abutments on the 
backs of the valve is then divided equally. See Sec. 73. 

70. To Determine The Requisite Length For Either The 
Steam-Valve Rod Or Stem Of A Duplex Pump proceed as 



58 



STEAM POWER PLANT AUXILIARIES 



[Div. 2 



follows: Place the valve-arm plumb (Fig. 68) and put the 
valve in its central position. The valve will be central when 
its outside edge at each end coincides with the outside edge 
of the corresponding steam port. If the valve stepn is 
missing, the valve-rod should be blocked up to a horizontal 
position. The length of the missing stem will then be given 
by the distance A (Fig. 68). In laying off this distance, the 
clearance, C, between the end of the stem and the wall of the 
steam-chest, should be greater than the steam-port width, 
assuming this to be equal to the maximum displacement of the 
valve from its central position. If the valve rod is missing, 
the stem should be inserted and the lost motion, L and L\ 
accurately adjusted. The length of the missing rod will then 
be given by the distance B (Fig. 68). 



■Valve Set Lrne-Anof- '; 
Line With Ports— 1 



Long 
Rocker 
Arm-, 




Fig. 68. — Method Of Finding Lengths Of Steam-Valve Stems And Rods Of 
Duplex Pumps. 



71. The Function Of The Valve-Stem Lost-Motion In 
Duplex Direct Acting Steam-Pumps (Fig. 63) is threefold: 
(1) It permits adjustment of the piston-stroke. (2) It causes a 
continuous piston-travel. (3) It prevents the pump from stopping 
in a position from which it cannot be started by admitting steam 
to the steam chest. Continuous piston-travel is secured by 
preventing simultaneous reversals of the piston-strokes. 
Assurance against a dead-center or non-starting position is due 
to the fact that when either steam-valve covers all four ports 
(Fig. 63) the opposite valve leaves an admission — and an ex- 
haust-port wide open. This feature of the duplex pump 
renders it well-adapted for periodic operation (Figs. 69 and 70) 
under governor control. 



Sec. 71] DIRECT-ACTING STEAM PUMPS 

fit"fP Supply. Pressure Pipe, 



59 



V- Throttle 
n Votive-, 




Main Pu'mp-^"^ 



■Auxiliary 'Duplex Pump 

Fig. 69. — Underwriters Fire-Pump Equipment For Connection To Sprinkler-Pipe 

System. 



Governor Pipe for Operation of Governor Under 

Gravit y Pressure of Water 

gfe^ j 1 1 i i i i i i 1 1 mini™ 



J|W-f fr 


% lira, 


III 1 1 


lilt; 


■ill 1 


llllllt: 


1 "!»!' 


~ni* 



Governor Pipe- for Operation of Gover- 
Under Discharge Water Pressure 




- -Suction 



\ 



Fig. 70. — Governor-Controlled Duplex Pump For Water- Service In Buildings. 



60 



STEAM POWER PLANT AUXILIARIES 



[Div. 2 



Explanation. — A duplex pump (Fig. 63) is presumed to have been 
correctly adjusted for running. The steam-pistons, Pi and P 2 , (Fig. 
71) are at mid-stroke. Likewise, the slide-valves, Vi and F 2 , are at mid- 
travel. Piston P 2 actuates valve Vi through the long rocker-arm, 



f" 




R 


Mmission 
, r -Ports 

y 


u 

4. 

; i 


Jl A 

Center ■ 
lines ; 
of-'.. 

Piston > 
Rools ■' 

f 7 / 


y 




i v* :: 


— 0fi\jMotion 




; 





Exhaust Por t— Admission Port—, 
Center Lines of 
Piston Roo/s^ 



Cross ! 
•-Afecvo/s , 




Centerlines of 
Votive Stems ■■ 



Fig. 71. — Pistons And Valves In Mid- 
Position. 




Aotmission Port— 
lost Motion- 



Fig. 72. — One Steam- Valve Shifted 
From Central Position. Piston Pi 
Starts Movement To Left. 



pivoted at Pi. Piston Pi actuates valve T 2 through the short rocker- 
arm, pivoted at P 2 . The rocker-arms stand perpendicularly to the center- 
lines of the cylinders. The lost motion of each valve-steam is equally 
divided between the ends of the valve. 

Valve 7 2 (Fig. 72) has been moved 
so that the steam-port S is open for 
admission of steam to the cylinder. 
It is presumed that this was done 
before the steam-chest covers (Fig. 
63) were put on. If the valves had 
been left as in Fig. 71, the pump could 
not start when steam would be ad- 
mitted through the throttle-valve in 
the supply-pipe. 

Steam has entered through port S 
(Fig. 72) and has driven piston P 2 in 
the direction of the arrow. This 
piston has moved just far enough to 
take up the lost-motion at the right- 
hand end of valve V\. Further move- 
ment will open steam-port Si. 
Piston P 2 (Fig. 73) has completed a stroke. Coincidentally, piston Pi 
has traveled far enough in the direction of the arrow to shift valve F 2 
to the mid-travel position, where it is on the edge of opening steam-port 
S 2 for the return-stroke of piston P 2 . 




•Aolmission- 
Ports 



Fig. 73.— Piston P 2 At End Of In- 
board Stroke And On Point Of Re- 
versal-Piston Pi Approaching End Of 
Inboard Stroke. 



Sec. 72] 



DIRECT-ACTING STEAM PUMPS 



61 



If the stem of valve V2 had less lost motion, the valve could not have 
been shifted (Fig. 72) far enough to give a full opening at port S. Hence 
valve V2 would have been moved to the mid-travel position (Fig. 73) 
before piston P 2 could have reached the end of the cylinder-bore. Thus a 
short-stroke would have resulted. On the other hand, if the lost motion 
were greater than the amount shown (Fig. 72), valve V2 could not reach 
mid-position (Fig. 73), and thereby close port S against admission of 
steam behind piston P 2 , coincident with the arrival of that piston at the 
end of the cylinder. 

Piston Pi (Fig. 74) has finished its stroke. Coincidentally, it has 
moved V 2 to the limit of its right-hand travel. Steam-port *S 2 has thus 
been fully opened for admitting steam behind piston P 2 . Piston P 2 
has, therefore, moved far enough on its return stroke to shift valve V\ 
to the edge of opening steam-port £3 for a reversal of piston P x . 



Exhaust Port---, Admission Port-. 

Center Lines of 
Piston Rods*. 




Exhaust Port- ■■ Admission Port- 



Fig. 74. — Piston Pi At End Of Inboard 
Stroke And On Point Of Reversal. Piston 
Pz Approaching End Of Head-End Stroke. 



Admission Port— ^ Exhaust Port— . 
Center Lines of ^¥ 
Piston Rods-. 




Center Lines of 
Valve Stems- ■ •' 



Exhaust Port—-' Admission Port— 



Fig. 75. — Piston P 2 At End Of Head- 
End Stroke And On Point Of Reversal. 
Piston Pi Approaching End Of Head-End 
Stroke. 



Piston P 2 (Fig. 75) has completed its return stroke. Coincidentally, 
piston Pi has traveled far enough on its return stroke to shift valve F 2 
to the edge of opening port S for another reversal of piston P 2 . 

Note. — Incorrect Adjustment Of The Valve-Stem Lost-Motion 
In Duplex Pumps May Be a Source Of Loss. When the pistons do 
not reach the limits of possible travel, they must make many more 
strokes than would otherwise be required to do the same amount of work. 
This means extra consumption of steam and cylinder oil, and extra wear, 
particularly of the water valves. 

72. The Points At Which The Cross-Heads Should Be 
Secured To Duplex-Pump Piston-Rods may be determined as 
follows: The packing should be removed from the piston-rod 



62 



STEAM POWER PLANT AUXILIARIES 



[Div. 2 



stuffing-boxes (Fig. 76) and the glands should be screwed up 
tightly. The cylinder-heads being removed, the steam- 
pistons should be pushed up solidly against the center-heads. 
A line, A, (Fig. 76) should then be scribed on each rod flush 
with the faces of the water-end glands or gland-nuts. The 
heads of the steam-cylinders should then be put on. The 



Steam- 
End-, 



Wetter, 
End >\ 




Where Center-Head 
Striking Point is Marked- 

Fig. 76. — Marking Center-Head Striking 
Point. 



Fig. 77. — Marking Cylinder-Head Strik- 
ing-Point. 



■Steam End Rocker Arm 



heads of the water-cylinders being removed, the steam-pistons 
should be pushed up solidly (Fig. 77) against the cylinder- 
heads. A line, B, (Fig. 77) should then be scribed on each 
rod flush with the faces of the steam-end glands or gland-nuts. 
The scribed lines, A and B, (Fig. 78) thus establish the striking- 
points. The lines should be prick-punched to make them 
discernible for future reference. By shifting the pistons until 

4 and B become equally dis- 
tant (Fig. 78) from the glands, 
the pistons will be placed ex- 
actly at mid-stroke. The 
crossheads should then be 
slipped along the rods until 
the rocker-arms (Fig. 78) 
stand plumb. The cross- 
heads may then be clamped 
to the rods. 




Cross head' 



Fig. 78. — Striking-Points Equally 
Spaced, Rocker-Arm Plumb, Crosshead In 
Correct Position. 



Note. — The crossheads of new duplex pumps are, generally, so secured 
to the rods as to preclude possibility of error in restoring the crossheads 
should they at any time be temporarily removed. The operation of 
finding the striking points (Sec. 72) is, however, necessary where the 
piston rods of an old pump have been renewed. 

73. The Correct Amount Of Valve-Stem Lost-Motion In 
Duplex-Pumps depends upon the service in which the pump 



Sec. 74] 



DIRECT-ACTING STEAM PUMPS 



63 



is to be used. Pumps designed to run at high speeds (Sec. 28) 
require considerably less valve-stem lost-motion than do 
pumps for slow-speed service. Generally, lost-motion (Fig. 
68), at each end of the valve, equal to about one-third of the 
admission-port width will suffice for ordinary service. 



Note. — The Valve-Stem Lost-Motion 
Duplex-Pumps, as those in boiler- 
feed and elevation service, should 
be such that each piston will travel 
nearly full stroke before shifting 
the opposite slide-valve to the 
admission edge of the steam-port. 



Valve Stem- 



3 



fe^ 



mm t aula 

Slio/e Valve - : Valve Seat-* 

Fig. 79.— Rigid Valve-Stem Con- 
nection Of Duplex-Pump Slide-Valve- 
Lost-Motion Provided Externally. 



In Slow-Running- 



Ao/ju sting Nuts 

■ -Sliding Block 




Fig. 80. — Mechanism For Outside Ad- 
justment Of Lost Motion In Duplex-Pump 
Valve Gear. 



Note. — The Valve-Stems Are Often Rigidly Attached (Fig. 79) 
to the slide-valves. In such cases a link mechanism (Fig. 80), with sliding 
blocks and tappets, is provided for adjusting the lost-motion outside the 
steam-chest. 



.■Cushion Valve ' 1/LTt** 



Slide Valve 




By Pass-- 



•Compression Spaoz 

Y . 

I- Cross-Section on Line X-Y E-loncj i t u ol i noil Section 

Fig. 81. — How Duplex-Pump Pistons Are Steam-Cushioned. 



74. Compression-Space In The Steam-Cylinders Of Duplex 
Pumps is the volume of cylinder-space S, (Fig. 81) in front 
of the piston, plus the volume of space in the admission-port, 
A, at the instant the piston has completely closed the corres- 
ponding exhaust-port E. 



64 



STEAM POWER PLANT AUXILIARIES 



[Div. 2 



Explanation. — The piston P, (Fig. 81) has reached a position in its 
travel wherein it prevents escape of steam through the exhaust-port 
E. Coincidentally the slide-valve, D, covers the admission-port A. 
Hence the piston will be cushioned in its further progress by compressing 
the steam ahead of it in the space S. 

Note. — Large duplex-pumps are equipped with cushion-valves, C, 
(Fig. 81) for adjusting the cushioning effect of steam in the compression 
spaces. This is done by controlling the flow of steam, through the by- 
pass, B, from the admission-port A to the exhaust-port E. 

75. The Relative Merits And Demerits Of Simplex And 
Duplex Pumps may be summarized as follows: (1) The flow 
of water in both the intake-and discharge-pipes of a simplex pump 



--Chest Piston 



.'-Slide Valve 

Stem of Auxiliary Valve which 
Admits Steam to Throw Chest 
Piston-P; 




\\\\\\\\\\\s\\^^^ 

Connection through which Chest Piston-P, ; Vibrating Rod which Actuates) 
is Operated by Chest Piston-P- - - ' Auxiliary Valve-Stem - 



Fig. 82. 



-Sectional Elevation Of Steam Cylinders Of A Burnham Compound Simplex 
Pump. 



must cease during piston-reversal. The water hammer which 
tends to result therefrom may, however, be prevented or 
modified by using (Sees. 61 and 65) air and vacuum chambers. 
(2) With a duplex pump (Sec. 71) the flow of water is prac- 
tically continuous. (3) The piston of a simplex pump travels 
the maximum set distance during- each stroke. The length of the 
stroke, after being fixed by adjustment of the auxiliary steam 
valve (Sec. 68), continues constant regardless of the retarding 
tendency of piston-, rod-, and cylinder-friction. (4) The 
pistons of a duplex pump may short-stroke. Short-stroking 
may be due to the retarding effect of friction between the 
pistons and cylinders and in the piston-rod stuffing-boxes. 



Sec. 76] 



DIRECT-ACTING STEAM PUMPS 



65 



(5) The simplex pump uses less steam for the same amount of 
work than does the duplex pump. This is due to smaller 
clearance spaces in the steam cylinder. 

Notes. — Simplex-Pumps Are Well-Adapted As Vacuum- and 
Air-Pumps in connection with surface-condensers. This is due to their 
comparatively small clearance spaces and immunity from short-stroking. 

Duplex-Pumps Are Well-Adapted For High-Pressure Service. 
They are also preferable where either a very high or a very slow velocity 
of flow is required. This is due to their practically continuous action. 

Fire-Insurance Underwriters require that Direct-Acting Steam 
Fire Pumps (Fig. 25) be of the duplex type. These pumps are com- 
monly connected to sprinkler-pipe fire-systems. In such cases auxiliary 
duplex pumps, A, (Fig. 69) are provided for making up the leakage from 
the sprinkler system and maintaining a constant pressure therein. 



Inlet from Junction Pipe-. 




High \ 
Pressure * 
Cylinder--' Exhaust Outlet to Feedwater" 

y Heotte^Contfenser or Atmosphere- ■ 

Fig. 83. — Sectional Elevation Of Steam Cylinders Of A Compound Duplex-Pump. 

Note. — Large Direct-Acting Steam Pumps Are Often Built 
With Compound Steam-Cylinders (Figs. 82 and 83). This is done to 
economize their steam consumption. See the Author's Steam Engines. 

76. The Steam-Piston Areas In Boiler-Feed Pumps 

(Sec. 28) are usually from about two to three times the water- 
piston areas. In boiler-feed' service the total head and the 
available steam-pressure are practically equal. A large 
excess of steam-piston area is, however, provided as a safety 
precaution. It conduces to prompt starting of the pump. 

77. Selection Of A Direct-Acting Steam Pump For Boiler- 
Feed Service is based upon two main factors: (1) The steam- 
ing capacity of the boilers to be fed. (2) A proper rate of 



66 STEAM POWER PLANT AUXILIARIES [Div. 2 

piston travel. The pump must be large enough to deliver, 
while running at a moderate speed, the maximum quantity 
of water that can be evaporated in the boilers. It is conven- 
tionally assumed that these conditions are fulfilled by selecting 
a pump that will deliver 45 lb. of water per hour per boiler 
horse power while running at one-half the rated normal speed 
of the pump. 

78. Pump Managment is discussed in the following notes. 
Although these directions are included in this Div. on Direct 
Acting Steam Pumps many of the suggestions apply with 
equal weight to pumps of any type. This material is quoted 
from the Coal Miner's Pocketbook: 

All Pumps, When New, Should Be Run Slowly until the parts have 
become thoroughly adjusted to their bearings, when the speed may be 
increased. Because a new pump works stiffly is no cause for alarm, for, 
while a machinist can properly construct the parts, he cannot always 
forsee the strains . caused by the action of the pump, when the parts are 
assembled and which require certain adjustments after the pump is at 
work. By running the pump slowly with the parts properly lubricated 
and making such adjustments as may be necessary, stiffness will gradu- 
ally disappear and the highest efficiency of the pump will then be at- 
tained, provided other matters on which the pump's action depend have 
received proper attention. 

The Causes That Affect A Pump, Impair Its Efficiency, And 
Prevent It From Performing Its Full Duty are: (1) wear; (2) the 
improper adjustment of valves, valve stems, and levers; (3) the improper 
packing of plungers and stuffing boxes; (4) drawing up the stuffing-box 
glands too tightly; (5) lost motion due to permitting the working parts to 
wear and not adjusting them to the new conditions; (6) accumulations of 
foreign matter under the valves or in the strainer; (7) broken valves and valve 
springs; (8) leakage in valves; (9) taking air in the suction pipe; (10) 
clogged or broken discharge pipes; and (11) the use of poor gaskets. 

Many Pumps Are Capable Of a Larger Capacity Than Is Ob- 
tained By The Low Speed At Which They Are Operated, but it is 
important that such pumps be run continuously, as any serious interrup- 
tion in pumping might cause trouble elsewhere. It is customary, there- 
fore, to keep on hand a supply of duplicate valves, moving parts, and 
packing, in order that when it becomes necessary to make repairs they 
may be made without great loss of time. 

A Common Cause Of Pumps Refusing To Work Properly Is Due 
To Their Taking Air Below The Suction Valves. Small leaks will 
cause the piston to jump owing to the water not entering through the 
suction valves soon enough to fill the entire chamber. This trouble 



Sec. 78] DIRECT-ACTING STEAM PUMPS 67 

may be remedied by making all joints in the suction pipe and between 
the pipe and the pump air-tight. Leaks may sometimes be detected by 
the hearing or by the flame from a candle being drawn toward the hole. 
If the leaks are small and not at the pipe joints, a coat of asphalt paint 
may stop them; if large, they should be drilled larger, the hole threaded, 
and a screw plug inserted. If the leak is at the joint between two pipes, 
the pipes should be uncoupled and screwed together again, using graphite 
pipe grease for a lubricant. Or, if the joint is a flanged one, a new gasket 
should be placed between the flanges, and the pipes lined up before the 
bolts are tightened. 

Sometimes, A Pump Fails To Catch The Water When Started 
Owing To Leakage Of The Valves In The Suction Chamber. The 
trouble may be caused by the valve and the valve seat being corroded; 
by chips or gravel getting under the valves and preventing them from 
seating properly; or by the valves and seats becoming worn so that leak- 
age cannot be prevented without changing the parts. 

Many Pumps Will Not Raise Water In The Suction Pipe When 
Empty, Owing To The Pump Having Been Idle For Some Time, 
but will continue to draw water after once being started. In such cases, 
it is necessary to prime the pump, by which is meant filling the suction 
pipe and part of the suction chamber, if there is one, and in some cases, 
also, the pump barrel, with water, so that the pump may start under 
conditions similar to those under which it must work. To prime the 
pump, open the cock, or valve, in the priming pipe and allow water 
from the column pipe to flow down into the suction pipe and the pump. 
When these are full, the valve is again closed and the pump is ready to 
start. 

Pumps Sometimes Fail To Raise Water When The Full Head Is 
Resting On The Valves In The Discharge Chamber. This may be 
due to air accumulating between the suction and the discharge decks, 
which air is compressed and expanded by the motion of the plunger. 
Air valves should be provided in the water cylinder to allow this confined 
air to escape. Violent jarring and trembling often occur if the discharge 
chamber is not provided with either an air chamber, where the lift is 
not above 150 ft., or with an alleviator, for lifts above that distance. 
This jarring is due to the column of water in the discharge pipe coming to 
rest suddenly between strokes and having to be again put in motion. 

In Case The Pump Column Is Filled With Water And The Pump 
Is Stopped, The Water Will Run Back Through The Pump If The 
Foot- Valve Is Not Tight. To prevent this, a gate valve or a check- 
valve is placed a short distance from the pump in the column pipe. A 
gate valve wears less than does a check-valve, and presents no obstruction 
to the flow of water when the valve is open. This valve is useful in the 
column pipe to maintain the pressure off the valves when the pump is not 
at work, and also for keeping water from running back into the pump 
chamber when the valves are being repaired. 

When Starting Compound Pumps, the steam pressure on the high- 



68 STEAM POWER PLANT AUXILIARIES [Div. 2 

pressure-cylinder piston is not always sufficiently powerful to move the 
plungers against the resistance of the water in the discharge pipe. But, 
by opening the gate valve in the by-pass piping, the pressure on the 
plungers is relieved for a sufficient number of strokes to allow the steam to 
reach the low-pressure piston, when the combined force of the two pistons 
will do the work. The by-pass pipe can then be closed. 

Valves In The Steam End Sometimes Wear Unevenly Or Their 
Stems, By Continual Action Wear And Cause, Lost Motion, thus 
causing a back pressure and irregular action. Anything wrong in the 
steam end can usually be determined by the irregular exhaust, but even 
this may be deceptive in case the water-end valves are leaking. If the 
steam valves are suspected, the steam chest cover may be raised for their 
inspection, but the valves should not be disturbed until it has been deter- 
mined, by moving the water piston backwards and forwards several 
times, that they do not open and close properly. The trouble may be in 
the levers or toggles that throw them. If so the correcting adjustments 
may be properly made without disturbing the valves. In many duplex 
pumps, there are very slight differences between the two sides, and the 
amount of the lost motion (Sees. 71 and 73) between the valve stem 
and the valve should be carefully adjusted. Too little lost motion will 
cause short stroking, while too much will allow the pistons to strike the 
heads. The adjustment requires skill. 

Sometimes, The Valve Seat Or The Valve Has Soft Spots That 
Wear Faster Than The Remainder Of The Valve And Seat. 
Through these slight depressions, steam will blow and cut both valve and 
seat if attention is not given them; back pressure will then seriously 
interfere with the working of the pump. If the defect is in the valve, a 
new one can take its place. But the valve seat, if a part of the steam 
cylinder, will require an entirely new cylinder, and hence it is economy to 
scrape the seat until the depressions are removed. A try plate made of 
steel having a perfectly level surface is covered with chalk and carefully 
rubbed over the valve seat. The elevations will have chalk on them, the 
depressions will not. The elevations are scraped with a chisel made of 
the best steel until they are worn down so that chalk sticks to every part 
of the seat alike. The valve is treated in the same way if it can be done 
without too much expense. The valve and the valve seat when remov- 
able should be sent to the shop to be reground. 

The First Step After A Pump Has Been Erected Is To Clean Out 
The Steam Piping. In order that this may be done without carrying 
foreign matter into the pump, the piping is left disconnected from the 
pump and steam at full boiler pressure is allowed to blow freely through 
the piping and valves for a few minutes. Steam is then shut off and the 
piping is connected to the pump. 

The Next Step Is To Blow Out The Steam Cylinders. To do this, 
the cylinder heads should be put on, leaving the pistons and valves out of 
the cylinders. The stuffing boxes should be closed, which is most 
conveniently done by placing a piece of board between the stuffing box 



Sec. 78] DIRECT-ACTING STEAM PUMPS 69 

and the reversed gland and then setting up the nut on the stuffing box 
studs. When the gland is drawn home by a nut outside of it, a circular 
piece of pine board may be placed between the end of the gland and the 
inside of the nut in order to close the opening through which the piston 
rod passes. Steam may now be turned on the main steam pipe leading to 
the pump; by opening the throttle valve wide at short intervals. 
Thereby the sand and scale, in the ports and other passages and spaces of 
the steam end, can be blown out. After the cylinders have been blown 
out, the heads and covers should be removed and all foreign matter 
blown into the corners and chambers of the cylinders removed by hand. 
The pistons, valves, cylinder heads, and other covers can then be put in 
place. The blowing out of the pipes and cylinders after erection is often 
neglected or but imperfectly done, with serious consequences to the machine. 
It cannot be too thoroughly done, particularly in pumps of the type in 
which the steam ports and exhaust ports are on top, for in this construc- 
tion the sand and grit are deposited in the bottom of the cylinder for 
the piston to ride on. 

The Packing Of All Rods And Stems Is The Next Step. If 
fibrous packing is used, the boxes should be filled full and the glands 
tightened down very moderately. The tightening of the glands can best 
be done when steam is on and the machine is in motion, when they 
should be tightened only sufficiently to stop leakage and no more. When 
excessive tightening is required to stop leakage, the packing should 
be completely renewed. Some pumps are fitted with metallic packing. 
This packing is usually prepared by specialists and fully guaranteed. 
Their directions for use should be carefully followed. In case of failure or 
unsatisfactory results, the makers should be consulted. 

The Oiling Of The Machinery Is The Next Step and is a very 
important one. All rubbing surfaces should be provided with suitable 
oiling devices designed for the particular place and service. The quality 
of oil should be carefully selected to suit the velocity and pressure of the 
rubbing surfaces on which it is used. For use within the steam cylinder, 
heavy mineral oil is the only oil capable of withstanding the high temper- 
ature. When starting up new pumps, only the best-quality oil should be 
employed, regardless of price. A liberal use of this oil for the first month 
will go far toward reducing subsequent oil bills. 

A Pump Must Often Run Continuously Without Interruption 
— For A Month Or Even Longer. This requires that all oiling devices 
be so arranged that they can be replenished and adjusted while the 
machine is in motion. It is a good plan to provide two sets of oiling 
systems for all of the principal journals. Then, if one fails the other can 
be used while the disabled one is being overhauled. All oil holes are 
generally stopped with wooden plugs or bits of waste twisted into the 
hole, or are otherwise protected while the machine is being erected. 
These should now be removed and the holes and oil channels thoroughly 
cleaned. Bearings should be flooded with oil at first to wash out any 
dust or grit that may have reached the rubbing surfaces. 



70 STEAM POWER PLANT AUXILIARIES [Div. 2 

The Steam End Is Now Ready To Be Warmed Up. (From now on 
the method of starting a pump is the same whether the pump is a new 
or an old one.) To warm up the steam end, the throttle is opened 
slightly and, with the drain cocks opened wide, steam is allowed to blow 
through the cylinder until no more water passes from the drain cocks. 
The steam by-pass pipes should be used where multiple-expansion 
pumps are being started. If the pump has a valve gear that can be 
operated by hand, the warming up can be hastened by working the valve 
back and forth slowly. While the steam end is warming up, the water 
end should be made ready by opening the stop-valve in the delivery pipe 
and otherwise insuring that the pump has a free delivery. If a stop- 
valve is fitted to the suction pipe, this should be opened. If the pump is 
compound or triple expansion, the water by-pass valves must be opened 
until the machine has made a sufficient number of strokes to bring the 
intermediate and low-pressure cylinders into action. Then the by-pass 
valves should be closed. If the pump is fitted with dash-relief valves, 
these should be closed before starting, keep the pistons as far from the 
heads as possible in starting. Should the pump exhaust into an inde- 
pendent condenser, this should be started and a vacuum obtained be- 
fore starting the pumps. 

To Start The Pump, the foregoing precautions having been observed, 
open the throttle slowly. Permit the pistons to work back and forth very 
slowly a few times, gradually increasing the velocity until full speed is 
attained. After the pump has been running a few minutes, close the 
drain cocks. If the pump has dash-relief valves, the length of stroke 
may now be carefully adjusted. 

To Stop The Pump, close the throttle, open the drain cocks, and (if 
there is one) close the gate valve in the discharge pipe. Finally shut 
down the condenser. If the pump is to remain stopped for some time, 
close the suction valve. 



79. The Causes Of Scoring Of Pump-Valve Stems and 
Piston Rods may be one, or all, of the three following: (l) 
Use of an improper packing, as a packing consisting of plain, 
unlubricated, hemp or rope fiber. (2) Permitting a fibrous 
packing to remain in the stuffing-box after it has become hard 
and brittle through age. When the packing attains this con- 
dition, attempts to prevent the steam from blowing out around 
the rod by drawing up on the gland will inevitably result in 
cutting and scoring the rod. (3) Use of an improper cylinder 
lubricant, as an oil containing an excess of animal fats. Such 
oils, in the presence of high temperature, evolve an acid which 
is particularly damaging to iron and steel. 



Sec. 80] 



DIRECT-ACTING STEAM PUMPS 



71 



80. Table Showing Duty And Steam Consumption Of 
Direct-Acting Pumps. Simple, Non-Condensing Steam 
Cylinder. (Values correct only for the typical efficiencies 
which are given. For other efficiencies modify values 
proportionately.) 



Non-jacketed, but lagged; wire drawing = 4.7 lb.; back pres 



16 lb. per sq. in. 



Boiler pressur 
Absolute initi 

M.e.p 

Card duty, m 




50 
60 
44 
45 


70 
80 
64 
50.5 


90 
100 
84 
53.5 


100 

110 

94 

55 


110 

120 

104 

56 


120 

130 

114 

57 


150 




160 


dlion ft.-lb 


144 
58.5 


Stroke, 
in. 


Mech. 

effic , 

per 

cent. 


Steam 

effic, 

per 

cent. 


Total 

effic, 

per 

cent. 


Actual duty, million ft.-lb. per 1,000 lb. 
dry steam = upper fig. Lb. dry steam 
used per water h.p. per hr. = lower fig. 


4 


55.0 


37.5 


9.5 
21 j208 


10.6 
187 


11.3 
175 


11.6 
171 


11.8 
168 


12.0 
165 


12.3 
161 


6 


65.0 


40.0 


1 11.7 
26 1169 

1 


13.1 
151 


13.9 
,43 


14.3 
139 


14.6 
136 


14.8 
134 


15.2 
130 


8 70.0 


42.5 


1 13.5 

30 147 

1 


15.2 
130 


16.1 
123 


16.5 
120 


16.8 
118 


17.1 
116 


17.6 
113 


10 


75.0 


45.0 


34 


15.5 
128 


17.2 
115 


18.2 
109 


18.7 
106 


19.1 
104 


19.4 
102 


19.9 
100 


12 


77.5 


47.5 


37 


16.6 
119 


18.7 
106 


19.8 
100 


20.4 
97 


20.7 
96 


21.0 
94 


21.7 
91 


15 


80.0 


50.0 


40 


18.0 
110 


20.2 
98 


21.5 
92 


22.0 
90 


22.5 
88 


23.0 

86 


23.5 
84 


18 


82.5 


52.5 


43 


19.4 
102 


21.7 
91 


23.0 

85 


23.7 
84 


24.0 
83 


24.5 
81 


25.2 
79 


24 


85.0 


55.0 


47 


21.0 
94 


23.7 
83.5 


25.1 
79 


25.9 
76 


26.5 
75 


26.9 
74 


27.5 

72 



72 



STEAM POWER PLANT AUXILIARIES 



[Div. 2 



81. Table Showing Duty And Steam Consumption Of 
Pumps. Compound, Non-condensing Steam Cylinder (See 
limitations in Table 80.) 



Non-jacketed, but lagged; wire drawing = 4.7 lb.; back press. = 16 lb. per sq. 



Boiler pressure 


50 


70 


90 


100 


110 


120 


150 


Absolute initial pi 






60 
1.94 


80 
2.24 


100 
2. 5 


110 
2. 62 


120 
2.74 


130 

2.85 


160 






3 16 


m.e.p. on area of h.p. cyl. . 




58.0 


88.5 


120.0 


136.0 


152.0 


168.8 


218.8 


Card duty, million ft. -lb. . 




60.0 


69.5 


76.5 


79.5 


82.0 


84.0 


89.0 




Mech. 


Steam 


Total 


Actual 


duty, 


million ft.-lb 


. per 


1,000 lb. dry 


Stroke, 


effic, 


effic, 


effic, 


steam = upper fig. 


Lb. dry steam per 


water 


in. 


per 


per 


per 


h.p. per hr. 


= lower fig. 










cent. 


cent. 


cent. 
























15.6 


18.1 


19.9 


20.7 


21.4 


21.8 


23.1 


6 


65.0 


40.0 


26 


127 


110 


99 


95 


92 


91 


85 










18.0 


20.8 


22.9 


23.8 


24.6 


25.2 


25.7 


8 


70.0 


42.0 


30 


110 


95 


86 


83 


80 


78 


74 










20.4 


23.6 


26.0 


27.0 


27.9 


28.5 


30.3 


10 


75.0 


45.0 


34 


97 


84 


76 


74 


71 


69 


65 










22.2 


25.7 


28.3 


29.4 


30.4 


31.1 


33.0 


12 


77.5 


47.5 


37 


89 


77 


70 


67 


65 


64 


60 










24.0 


27.8 


30.6 


31.8 


32.8 


33.6 


35.6 


15 


80.0 


50.0 


40 


83 


71 


65 


62 


60 


59 


56 










25.8 


29.9 


32.9 


34.2 


35.3 


36.1 


38.3 


18 


82.5 


52.5 


43 


77 


66 


60 


58 


56 


55 


52 










28.2 


32.6 


36.0 


37.4 


38.4 


39.5 


41.9 


24 


85.0 


55.0 


47 


70 


61 


55 


53 


52 


50 


48 










30.0 


34.0 


38.2 


39.7 


41.0 


42.0 


44.5 


36 


87.5 


57. 5 


50 


66 


58 


52 


50 


48 


47 


45 



82. Table Showing Duty And Steam Consumption Of 
Pumps. Compound, Condensing, Steam Cylinder. (See 
limitations in Table 80.) 



Sec. 82] 



DIRECT-ACTING STEAM PUMPS 



73 



L-p. cyl. jacketed and lagged; wire drawing = 4.7 lb.; back press. = 6 lb. per sq. in, 



Boiler pressu 
Absolute init 
Ratio of cyls 


re 






70 
80 

3.65 
116.2 
91.0 


90 

100 

4 

151 

96.5 


100 
110 
4 
168.5 

98 


120 
130 
4 
203.5 
101.5 


150 
160 
4 
256 
104.5 


170 

180 

4 

291 


180 
190 








4 








308. 5 


Card duty, million ft.- 


lb 




106.5 107 


Stroke, 
in. 


Mech. 

effic, 

per 

cent. 


Steam 

effic, 

per 

cent. 


Total 

effic, 

per 

cent. 


Actual duty, million ft.-lb. per 1,000 lb. 
dry steam = upper fig. Lb. dry steam per 
water h.p. per hr. = lower fig. 


10 


75.0 


55.0 


41 


37.4 
53 


39.6 
50 


40.2 
49 


41.6 
48 


42.9 
46 


43.7 
45 


44.0 
45 


12 


77.5 


57.5 


45 


41.0 
48 


43.4 
45 


44.1 
45 


45.5 
44 


47.1 
42 


48.0 
42 


48.0 
41 


15 


80.0 


60.0 


48 


43.7 
45 


46.4 
43 


47.0 
42 


48.7 
41 


50.1 
40 


51.1 
39 


51.3 
38 


18 


82.5 


62.5 


52 


47.4 
42 


50.0 
40 


51.0 
39 


52.8 
37 


54.3 
37 


55.4 
36 


55.8 
35 


24 


85.0 


65.0 


55 


50.0 
40 


53.0 
37 


54.0 
37 


55.9 
35 


57.6 
35 


58.6 
34 


59.0 
33 


36 


87.5 


67.5 


59 


53.8 
37 


57.0 
35 


58.0 
34 


60.0 
33 


61.7 
32 


62.8 
32 


63.1 
31 


48 


90.0 


70.0 


63 


57.2 
35 


60.8 
33 


61.9 
32 


64.0 
31 


65.8 
30 


67.1 
30 


67.4 
29 



QUESTIONS ON DIVISION 2 



1. What is a double-acting suction pump? 

2. Explain the operation of a double-acting suction pump. 

3. What velocity of water-flow is recommended for the suction-piping of steam 
pumps? For the discharge-pipe of a simplex pump? For the discharge-pipe of a 
duplex pump? 

4. What is a piston-pump? A plunger-pump? 

5. What is an outside-packed plunger-pump? An inside-packed plunger-pump? 

6. What is the distinction between an outside-end-packed plunger pump and an 
outside-center-packed plunger pump? 

7. For what maximum discharge-pressures are piston and plunger pumps respectively 
adapted? 

8. Explain the method of packing a water-piston with hydraulic packing. 

9. In what class of pump service are soft rubber-composition valve discs especially 
suitable? In what class of pump service are metal valve discs especially required? 

10. How are the water-valves arranged, with reference to the pistons or plungers, in 
horizontal direct-acting steam-pumps? Which arrangement is commonly used in 



74 STEAM POWER PLANT AUXILIARIES [Div. 2 

pumps for high-pressure service? Which arrangement is recommended for vacuum 
pumps? 

11. What is the function of an air-chamber? 

12. Explain the operation an of air-chamber. 

13. Why are air chambers less necessary on duplex pumps than on simplex pumps? 

14. What is the highest level, consistent with good service, to which the water may 
rise in an air chamber? 

15. What is a snifter, as used in air-chamber service? How does it work? 

16. Describe a method of recharging air-chambers in pumping systems working under 
pressures up about 1,000 lb. per sq. in. 

17. What is the proper ratio of air-chamber volume to water-piston displacement in a 
single pump? In a duplex pump? In fire-pumps? 

18. What is the function of a vacuum chamber? 

19. Explain the operation of a vacuum chamber. 

20. What is a simplex steam-pump? A duplex steam-pump? 

21. What is meant by the term steam-thrown, as applied to the steam-valves of 
simplex pumps? 

22. Upon what adjustment does the length of stroke of simplex pumps with external 
valve gears commonly depend? 

23. Assuming that the crossheads are properly secured to the piston-rods, what three 
principal adjustments are necessary for correctly setting the steam- valves of a direct- 
acting duplex pump? 

24. What is the three-fold function of the valve-stem lost-motion in duplex direct- 
acting steam pumps? 

25. Describe the cycle of steam-valve motion in the operation of a duplex pump. 

26. What disadvantage ordinarily results from incorrect adjustment of the valve- 
stem lost-motion in duplex pumps? 

27. Describe a method of marking the striking-points of duplex-pump pistons. 

28. How much lost-motion should the steam valve-stems of duplex pumps ordinarily 
have? 

29. What is meant by compression-space in the steam-cylinders of duplex pumps? 

30. What are cushion-valves on duplex pumps? 

31. What are the advantages of simplex steam-pumps as compared with duplex 
steam-pumps? What are the disadvantages? Which type is recommended for fire- 
protection service in buildings? Which type is recommended for use in connection 
with surface condensers? Why? 

32. What considerations govern the proportioning of water-piston areas to steam- 
piston areas in boiler feed pumps? What is the usual proportion? 

33. What two principal considerations govern the selection of a direct steam driven 
boiler feed pump? 

34. What are the causes which may impair the effectiveness of a pump when it is in 
service? 

35. Explain some conditions which may cause a pump to fail to raise water. Give 
remedies for each. 

36. Explain the method of repairing the steam valve and valve seat in a pump 
when they are badly worn. 

37. Enumerate and explain the successive steps in erecting a pump. 

38. Discuss steam-pump lubrication. 

39. Explain how a pump should be started. 

40. What are the steps in stopping a pump? 

PROBLEMS ON DIVISION 2 

1. A direct-acting steam-pump for low-speed service has a plunger diameter of 12 in. 
The plunger is inside-packed. The plunger-rod is of 3-in. diameter. How many 
flat disc valves, each of 4-in. diameter and 0.25-in. lift, should there be in each set of 
suction and delivery valves in this pump? 



DIVISION 3 

CRANK-ACTION PUMPS 

83. Crank -Action Pumps include piston or plunger pumps 
of all forms which depend for their operation on the circular 
motion of a crank-shaft. They may be classified as follows: 
(1) Crank-and-fly-wheel pumps in which the reciprocating move- 
ment of the pump piston or plunger is derived directly (Figs. 



Air Chamber- ^ 

.-Steam Cylinder 

Vacuum Chamber 
Steam Mounted At 

Supply^ Suction Inlet- 



Discharge Nozzles 



Fly Wheel- 




Eccentric <' 



Fig. 84. — Steam-Driven Crank-And-Fly- Wheel Pump. 

84 and 85) or indirectly (Fig. 86) from the reciprocating 
movement of a piston in a steam cylinder but is dependent 
for its continuance upon the inertia effect of the rotative move- 
ment of a crank-shaft and fly-wheel. (2) Crank-action power 
pumps in which the reciprocating movement of the pump 
piston or plunger is derived from the rotative movement of a 
mechanically-driven crank-shaft. Figures 87, 88 and 89 show 

75 



76 



STEAM POWER PLANT AUXILIARIES 



fDiv. 3 



belt driven power pump and Fig. 90 shows a gear driven power 
pump. 




Fig. 85. — Horizontal Section Of A Double-Acting Duplex Crank-And-Fly- Wheel 

Pump. 



Connecting 




Fig. 86. — Crank-Action Pump Of The Walking-Beam Type (Pumps Of This Type Are 
Now Practically Obsolete). 



84. In The Operation Of Crank-And-Fly-Wheel Pumps 

the steam is worked expansively in the driving cylinders 
instead of being admitted during the entire stroke of the piston 
(Sec. 90), as in the operation of direct-acting steam-pumps. 



Sec. 84] 



CRANK-ACTION PUMPS 



77 



Hence, a fly-wheel is necessary to insure approximately uni- 
form movement throughout the stroke. (See the author's 
Steam Engines.) The pump piston or plunger is usually 
connected directly to the piston-rod (Fig. 91) of the driving 
cylinder. Hence, the function of the crank-and-fly-wheel is 
only to insure minimum variation of the rotative speed. 



Tight Pulley. 



Disc-. 
Crank 





Fig. 87. — A Belt-Driven Single-Acting 
Pump For Boiler-Feeding. 



Fig. 88. — Combination High-Service And 
Low-Service Belt-Driven Pumps. 



Crank-and-fly-wheel pumps are generally more economical 
than direct-acting steam pumps. This is due to the expansive 
use of steam in the cylinders and to the better valve action 
which is obtained, as in the steam engine, by the use of 
properly-designed Corliss and slide valve-gears. Hence, they 
are chiefly employed where steam-driven pumps are desired 
but considerations of economy preclude the application of the 
direct-acting type. 



78 



STEAM POWER PLANT AUXILIARIES 



[Div. 3 



Note. — An Advantage Claimed For Crank Action As Compared 
With Direct Action in the operation of steam-pumps is that crank- 



Vtght 

Pulley 




Discharge Valves- 



Fig. 89.— A Belt-Driven Double-Acting Fig. 90.— Belt Driven Single-Acting Plunger 
Pump For Boiler Feeding. Pump For Boiler Feeding. 



.-Flywheel 




Suction 
Inlet- ■■■ 



Fig. 91.— Single- Acting Crank-And-Fly wheel Pump For Hydraulic Elevator Service. 

action entirely obviates the short-stroking of the pistons (Sec. 75) which 
is liable to occur with direct-acting pumps. Also, since the limits of 



Sec. 85] 



CRANK-ACTION PUMPS 



79 



the piston stroke are definitely fixed, less clearance is necessary at the 
ends of the cylinders. Crank action, as a rule, permits of a higher piston 
speed than is practicable with direct-action. This is due to the energy 
which is stored up in the moving mass of the fly-wheel at the termination 
of the stroke. This energy is available for reversing the motion of the 
piston. With direct-action, the reversal of the stroke is effected solely 
by steam pressure. 

Note. — Steam-Driven Pumps Of The Crank-And-Fly-Wheel Type 
Were Formerly Extensively Used In City Water Works and large 
hydraulic elevator installations (Fig. 91). The comparatively large units 
designed for this class of service are called pumping engines. Pumps (Fig. 
84) which are used in sugar mills for pumping molasses are also of this 
type. 



AnrEcczntric-- 



rPoppef Discharge Valves? 




Fig. 92. — Alberger Rotative-Reciprocating Dry- Vacuum Pump. 



Note. — A Majority Of Steam Air-Compressors And Dry-Vacuum 
Pumps are, strictly speaking, included in this group but are more con- 
veniently discussed under other headings. For vacuum pumps (Fig. 
92) see Sec. 354. 

85. The Steam Consumptions Of Crank-And-Fly-Wheel 

Pumps are determined by the same general factors that govern 
the steam consumptions of steam engines in similar classes of 
service. These factors are, mainly, the type of steam valve 
gear that is used and the methods of operation — whether sim- 
ple or compound, condensing or non-condensing. Slow-speed 
crank-and-fly-wheel pumps with single steam cylinders of the 
simple slide-valve type consume, when operated non-conden- 



80 



STEAM POWER PLANT AUXILIARIES 



[Dry. 3 



sing, about 50 lb. of steam per indicated horse power hour. 
High-duty crank-and-fly-wheel pumps with compound steam 
cylinders and Corliss steam valves consume, when operated 
non-condensing, about 25 lb. of steam per indicated horse 
power hour. With condensing operation, the steam consump- 
tion of these high-duty pumps may be as low as 10 lb. of steam 
per indicated horse power hour. 

86. The Advantages And Disadvantages Of Crank-And- 
Fly-Wheel Pumps in comparison with direct-acting steam- 
pumps may be enumerated as follows: (1) Steam-consumption 



Crank Shot ft-. 



Main Gear-. 




J _Jj ':, Plungers 



Discharge Outlet- 

Fig. 93. — Triplex Pump For Heavy Liquids. 



is generally more economical. (2) May be run at higher speeds 
for most classes of service. (3) First cost is greater. (4) 
Require greater operating attendance. (5) Cost of maintenance 
is greater. 

Note. — The water-ends of crank-action pumps are built in many- 
respects like the water-ends of direct-acting pumps, which are discussed 
in the preceding Div. The information there given relative to the care 
of valves, packing, and management in general applies here to pistons, 
glands, plungers, and other parts. The subjects of piping, pressures, 
heads, suction and the like are also largely omitted here as they are dis- 
cussed in Divs. 1 and 2. 



Sec. 86] 



CRANK-ACTION PUMPS 

MainGMr--...^ ^eSS^^m*.. p} n j on _. 



81 




£■'■■■ 


■ssk 




_JJS 


4- 


::■; 




s 


■ 




£ 


= 


i ''! 


N 


■: 




<: 






<5 


: ii"' 


. 'Ill: 


\ 




lllliill 



'•Discharge Valves' " Suction Inlet 
Fig. 94. — Sectional View Of Single-Acting Triplex Pump. 



Crank 
Shaft-- ..J 

i, 



Mam fear-, 




Fig. 95. — Sectional View Of Double-Acting Triplex Pump. 



82 STEAM POWER PLANT AUXILIARIES [Div. 3 

87. Crank-Action Power Pumps may be divided into three 
main classes: (1) Simplex power pumps (Fig. 87) in which the 
pumping operation is performed by a single piston or plunger, 



Spur Czar On Crank Shaft, 
Cranky. ^^j?£^L 

\ ^Vt Motor-. 




Fig. 96. — Pump Driven By Motor Through Spur Gearing. 

rCrunk Shaft Pulley 

■-Pump 




Fig. 97. — A Belt-Driven Power-Pump. 



Crank Shaft 
^Sprocket Wheel Pump % 

-Crank \ JZII [j 

je-Zha'm 

Motor, 




Fig. 98. — A Chain-Driven Power Pump. 

(2) Duplex power pumps (Fig. 90, 200, and 201) in which the 
pumping operation is performed by two pistons or plungers 
operated by a common crank-shaft, (3) Triplex power pumps 



Sec. 87] 



CRANK-ACTION PUMPS 



83 



(Fig. 93) in which the pumping operation is performed by 
three pistons or plungers operated by a common crank-shaft. 
These pumps may all be single acting (Fig. 94) or double 
acting (Fig. 95). If the pump is double acting, the plunger 
may be in two parts as in Fig. 53. 




Fig. 99. — "Goulds" Triplex Deep Open- Well Pump. 



Note. — Power may be supplied to power pumps by electric motors 
(Fig. 96), gas or gasoline engines, water-wheels, steam engines or line-shaft- 
ing variously driven. This power may be transmitted to the crank-shafts 
of the pumps by means of belts (Fig. 97), chains (Fig. 98), gears or rope- 
drives. The pump crank-shafts may also be connected directly to the 
drive-shafts of the prime movers. 



84 



STEAM POWER PLANT AUXILIARIES 



[Div. 3 



88. Crank-Action Power Pumps Are Designed And 
Arranged In Various Ways For Deep-Well Service. — Since 
wells are frequently more than 22 feet (practical suction lift, 
Sec. 2) deep, it is often necessary to install pumps with their 
cylinders below the ground level so as to force the water out 




Fig. 100. — A Motor-Driven Deep-Well Pump. 



by pressure. Sometimes wells have large sectional areas and 
are comparatively shallow. For such, the under-ground por- 
tions of the pumps may be installed (Fig. 99) very much like 
ordinary power pumps. They are, however, provided with 
elongated plunger-rods which connect to the crank-shafts 



Sec. 891 



CRANK-ACTION PUMPS 



85 



Plunger Drop 
Rod—>Tl Pipe) 



located above ground. More often deep wells are merely 
drilled holes ranging possibly from 2 inches to 
12 inches in diameter protected by metal-tube 
casings. They may be several hundred feet 
deep. For such wells it is necessary to use 
the so-called deep-well or artesian-well pumps 
(Fig. 100) which have been especially de- 
signed for this service. 
V 1 89. Crank-Action Pumps For Deep-Well 

-^ Service are of three kinds: (1) Single-acting 

pumps discharging on the up-stroke only 
(For exception see Sec. 90), Fig. 101. (2) 



Driving-Rod . 
Connection--' 

Brass Ball-Valve- 
For Down-Stroke 



Brass Cylinder 
Shell--... 



■Drop Pipe 

Rubber Disc-Valve 
For Up-Stroke 
Discharge-, 



Flange Leather 
Packing 



Rubber Disc-Valve 
For Down-Stroke ,■ 
Suction '" 



Ducts Connecting 
Outer And Inner 
Plunger-Tubes- . . 



Rubber Disc-Valve 
For Up-Stroke Suction 




Cup-Leather Plun- 
ger Packing. 



Flange Leather 
Packing. 



■Hollow Tail-Rod 



Fig. 101.— Cylinder 
Of Single- Acting Deep- 
Well Pump. 



Fig. 102.— Cylinder Of Double-Acting Deep- Well 
Pump (Plunger Making An Upward Stroke). 



Double-acting pumps having one plunger but discharging on 
both the up-stroke and the down-stroke, Fig. 102. (3) Two- 



86 



STEAM POWER PLANT AUXILIARIES 



[Div. 3 



stroke pumps having two plungers 
operating in one cylinder controlled 
by two well-rods, Figs. 103 and 104. 
Pumps of this last type discharge 
almost continuously and are fre- 
quently used in deep-well service. 

Note. — Some Engineers Prefer An 
Air-Lift for certain deep-well pumping 
applications, because it has no moving 
members (except the compressor), is inex- 
pensive, and has no parts requiring repair 
underground where they are inaccessible. 
They are not as efficient from a power 






-CroinkGmrs 



Puliey-M 



Frame 




■ So/t'cf , 
- Suction 
Root-- 


"Hi 






Gui'ofe Roofs--- 


BfrrfHt: 1 



Discharge Pipe 



Fia. 103. — Chippewa Power-Driven Deep-Well- 
Pump Head Or Operating Gear. 




Uc'£ 





L £z 







m 







Fig. 104. — Deep- Well Pump 
Cylinder Fitted With Differ- 
entially-Operating Plungers. 



Sec. 90] 



CRANK-ACTION PUMPS 



87 



standpoint as pumps but are proof against damage by grit and are not 
likely to get out of order. 

Explanation. — Fig. 100 shows a typical motor-driven deep-well in- 
stallation. It may be single-acting if used with the cylinder and plunger 
of Fig. 101 or double-acting if used with the c}dinder and plunger of 
Fig. 102. The lower ball-valve (Fig. 101) opens on the up-stroke allow- 
ing the pump to fill with water. On the down-stroke, the lower valve 
seats and the upper valve opens allowing the water in the cylinder to 
flow past the plunger. On the next up-stroke, the water is lifted up the 
drop-pipe. The double-acting plunger (Fig. 102) operates similarly to 
the single-acting plunger on the up-stroke. On the down-stroke, how- 
ever, the water, instead of merely passing the plunger, is forced up the 
drop-pipe through the hollow plunger-rod. Meanwhile more water is 
drawn into the upper part of the cylinder through the hollow tail-rod. 

90. A Compound Or Two-Stroke Deep-Well Pump Operat- 
ing Gear is shown in Fig. 103. Its plungers and cylinder, 
which are located underground, are sim- 
ilar to those shown in Fig. 104. The two- 
stroke type of pump has the advantage 
over the single-acting type that it insures 
a more nearly continuous movement of 
vertical- water column. Its advantage 
over the single-plunger type is that the 
two plungers are of about the same 
weight and balance each other; as one is 
going up, the other is coming down. 



Explanation. — As the geared cranks A and B 
(Fig. 103) revolve, one or the other of the two 
plungers L and T (Fig. 104) is on the up-stroke 
continuously, except at dead-center. When the 
plunger T is on the up-stroke, its valve Vi seats 
and water is forced by it up the drop-pipe, while 
valve V 2 (Fig. 105) opens and allows water to 
pass plunger L which is then on the down-stroke. 
On the return stroke, the valve V 2 seats and 




'-Three Cup Washers' 
Fig. 105. — Cross Section 

water is forced through valve V x and on up the ° f Pump Plun e er Shown 
pipe. InFig " 104 - 

Note. — Most deep-well pump plungers are packed with leather cup- 
washers (Fig. 106). The plunger rods at the top of the drop pipes are 
packed, usually, with fibrous packing in the same way as are piston-rod 
glands. The valves used are either ball-valves (Fig. 101), disk valves or 
conical seated valves (Fig. 107). Plunger-rods or well-rods of the single- 
acting type are usually of wood with steel fittings and should be fitted 



88 



STEAM POWER PLANT AUXILIARIES 



[Div. 3 



with guide-couplings (Fig. 108) which slide on the inside of the drop- 
pipes and prevent the rods from buckling. Well-rods for double-acting 
pumps are generally made of wrought iron pipe, on account of the com- 



Eot&e Chamfered-. 





Fig. 106. — Leather Cup For Packing 
A Deep- Well Pump Plunger. 



Fig. 107. 



Water 
Passage 



-Plunger-Valve For Deep- 
Well Pump. 



pression strain on the down stroke. Guide-couplings should be used 
about every twenty feet. Two-stroke pumps have a solid steel or iron 
rod (Fig. 109) driving the lower plunger (Fig. 110). This rod slides 




Fig. 108. — Steel Guide Coupling For Well-Rods Of Deep- Well Pumps. 

inside of a hollow tube or pipe which drives the upper plunger (Fig. 111). 
Both rods must be guided and packed. In open-well pump installations, 
the plunger rods are guided (Fig. 99) with grooved rollers. 



• ■ -Threaded Shank 
..-■■Threaded End 



r. k 

" ■ Sec tion Of Solid Con nee Una ■ Rod > 

3 Couplings-; 

Section Of Hollow Connecting-Rod-., \ 



m 



Fig. 109. — Connecting Rods For Operating Plungers Of Two-Stroke Deep- Well Pump. 



91. The Characteristics Of Crank -Action Pumps are very 
different from those of pumps of the direct-acting type. Com- 
pare the indicator diagrams for the steam- and water-ends of 



Sec. 91] 



CRANK-ACTION PUMPS 



89 



the crank-and-fly-wheel pump shown in Fig. 112 with cor- 
responding ones for direct-acting pumps shown in Figs. 22 
and 23. The difference between the diagram for the steam- 
end in Fig. 112 and that in Fig. 22 is due to cut-off at about one 
third stroke in the crank-action pump and non-expansive 
use of steam in the direct-acting pump. The difference 
between the water-end diagram shown in Fig. 112 and that 
shown in Fig. 23 is due partly to the more rapid movement of 



Solid 
•Suction 
Roof 



Lift Of 
Valve-, 





Fig. 110. — Lower Plunger Of Two-Stroke 
Deep- Well Pump. 



Fig. 111. 



-Upper Plunger Of Two-Stroke 
Deep- Well Pump. 



the piston in mid-stroke in the crank-action pump and uniform 
movement throughout the stroke in the direct-acting pump. 
The type of water-end diagram of Fig. 112 is characteristic 
only for low-pressure and high-speed crank-and-fly-wheel 
and power pumps. Higher pressures and lower speeds in 
crank-action pumps produce indicator diagrams which are 
more nearly rectangular. The higher-speed water-end dia- 
grams are characterized by sharp pressure peaks and very 
irregular pressure curves. 



90 



STEAM POWER PLANT A UXILI ARIES 



[Div. 3 




A tmospheric L me~* 
Stroke Inches (Reduced) 




Fig. 112. — Steam-End And Water-End Indicator Diagrams For Small Low-Pressure 
Crank-And-Fly- Wheel Pump. 



' Graph Showing Intermittent Suction--, 
g * Graph Showing In+ermittent Discharge-^ 



8% 



- — -J--T-- -J1 


% + -F~^^ ~t u 


s / V 


\ it 2 \ t 


' ' V i . , / \ ■ 


\No Discharge Line-f.^ V 


_ _, L -r-£ = 






S^ J X 


it s ^ s^ 


^-^ ^ 


90° 160 J 770 1 560 90" BO" 



Angular Position Of Crank 



Fig. 113. — Graph Showing Rates Of Suction And Discharge Of A Simplex Single- 

Acting Pump. 







Line, Of No Suction And No Discharge 
Point Of No Suction And NoDischarge- 








•ji 








+ 


_ I . I i -I i I L_ 




















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cs- 




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K 






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k 


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r 






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3£ 










<l 


fr 


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s 


c- tf 


■%- 












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w .2 










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K 






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v 












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JRctfz OfSuctio 
I I I 


n: 


























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9 


o" 




ia 


0" 




2 


0" 




3t 


0" 




9 


T 




\( 


J 



•^Angular Position Of Cranks 

Fig. 114. — Graph Showing Rates Of Suction And Discharge Of The Individual 
Cylinders Of A Duplex Single-Acting Pump With Cranks 180 deg. Apart Or Of A 
Simplex Double-Acting Pump. 



Sec. 92] 



CRANK-ACTION PUMPS 



91 



92. The Rate Of Suction And Discharge Graphs for a sim- 
plex single-acting crank-action pump are shown in Fig. 113. 
The lines of no-discharge and no-suction are separated to show 
the intermittent nature of the action of pumps of this type. 







*fi 


Position Shown In Fig. 117 ,-Qraph Of Total Discharge Porte Of Pump 
6Ct 90* 120* 150' 180*/ 210" 240* Z70* TO* TO' %0* %' 60* 


90* 


CI 








s 








v 1 












Y 


























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u 


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ity/e /-o/- 




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r~ 


Discharge-, 




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-Ab^e For 
















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k 


Ab'/e /cv 






































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Cylinder^ 


p.. 




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Co 








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V 










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x | 




















































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q 


4 

UK 1 


ar P 


OS 


It 

t 


5 s 

or 


IS 


Of 


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C 


5^ 


E 
nk 


5" 

5 


I 


.5' 


Z 


55" 


2 


35* 


3 


5" 


* 


5' 


1 


5' 


4 


y 


h' 


A 



Spur Wheel 



Fig. 115. — Graph Rates Of Suction And Discharge Of The Individual Cylinders {A, 
B and C, Fig, 117) Of A Single-Acting Triplex Pump. Also The Resultant Or Total 
Discharge Of All Of The Cylinders. 

Fig. 114 shows graphically the suction and discharge rates for 
a single-acting duplex pump with cranks 180 deg. apart. A 
double-acting simplex pump has the same characteristics. 
Pumps of these types have instants of inaction at the ends of 
the strokes as shown at A on the graphs. 

Explanation. — The Suction And Discharge Graphs For A Tri- 
plex Pump similar to the one shown 
in Fig. 93 are shown in Fig. 115. A 
pump of this type has a crank-shaft 
(Fig. 116) having three cranks which 
are set 120 deg. apart. Fig. 117 shows 
diagrammatically the position of each 
crank separately at a given instant. 
The graphs in Fig. 115 show the rates 
of discharge and suction of each of 
the individual cylinders A, B, and C 
(Fig. 117) and also of the whole pump. 
The line XX' (Fig. 115) represents the position of the plungers at the 
instant considered in Fig. 117. The distances of the points A', B', and C 
from the line of zero suction and zero discharge (Fig. 115) represent the 
rates at which the cylinders A, B, and C are sucking or discharging 
at the instant considered. Cylinder A is at dead-center and, therefore, 



•Crank Shaft 
\ A 




-Cranks 



Fig. 116. — Main Gear And Crank- 
Shaft Of A Triplex Power Pump. 



92 



STEAM POWER PLANT AUXILIARIES 



[Dw. 3 



point A' is on the zero discharge line. Cylinder B is nearing its maximum 
rate of discharge as shown by the rise of graph B at B'. Cylinder C 
has passed its maximum rate of suction, as shown by the upward slope 
of the graph C at C". Graph Y represents the total discharge rate and 
graph Z the total suction rate of the pump. 

.i 70 " .-Crank-. l9C ? Direction Of Rotation-, 




Fig. 117. — Diagrammatic Illustration Of A Single-Acting Triplex Pump, Showing 
Relative Positions Of Its Elements At A Given Instant. (See preceding illustration 
for graphs.) 

93. The Allowable Speed For Crank-Action Pumps varies 
over a wide range according to conditions. The following 
values are from various sources: 

94. Table Showing Typical Crank -Action-Pump Piston 
Speeds. 



Type of pump 



Piston speed, 

feet 

per minute 



60 inch stroke crank-and-fly-wheel pump 

30 inch stroke crank-and-fly-wheel pump 

15 inch stroke crank-and-fly-wheel pump 

18 inch stroke geared power water pump 

Deep-well pumps about 24 inch stroke 

Water supply pumps 5" X 12" to 9" X 16" 50 lb. to 

1000 lb. pressure 

Water supply pumps 5" X 12" to 9" X 16" up to 

3000 lb. pressure 

Hydraulic pumps up to 5000 lb. pressure 



300 
250 
200 
100 
100 

100 

80 
50 



Note. — For thick liquids and high suction lifts the allowable piston 
speeds are lower than specified above. 



Sec. 95] 



CRANK-ACTION PUMPS 



93 



95. Selection Of Pumps For Liquids Other Than Water 

(Marks' Handbook) should be discussed usually with the 
pump manufacturers. The following indicates usual practice: 



Liquid 


Material 


Liquid 


Material 


Brine 

Caustic 

Hydrochloric acid 


Brass fitted 
All iron 
Lead lined 


Oil 
Sewage 


Brass fitted 
Brass fitted 
Large openings 



Note. — Corrosive Liquids are handled ordinarily by air pressure or 
in properly-lined centrifugal pumps. Gummy liquids are handled pref- 
erably in pumps with large ball-valves. Volatile non-corrosive liquids, 
such as alcohol and gasoline, may be handled the same as water except 
that the liquid must always flow to the pump by gravity. 

96. Selection Of Proper Pump Power And Capacity is a 

matter of computation, as explained in Div. 1, but the follow- 
ing table of typical pump data shows, in a general way, the 
size and power necessary for a given capacity. 

97. Table of Typical Crank -Action Pump Data. 









Power re- 






Type 


Bore 
and 


Speed, 


quired in h.p. 
per 100 lb. 


Pulley 


Capacity, 




stroke, 
inches 


r.p.m. 


per sq. in. 
head 


size, in. 


gal./min. 


a o 


2X2 


65 


*0.24 


12 X in 


*3.4 


o ^ .2 

'43 o -g 


3X4 


55 


*1.00 


14 X 3 


*13 


§ « * * 


4X6 


55 


*3.11 


18 x zy 2 


*35 


ingle- 
duple 
ouble 
simp] 


6X8 


50 


*6.40 


20 X 5 


*95 


4 X 12 


42 


*3.96 


is x zy 2 


*54 


Ul Q 


10 X 12 


40 


*23 . 00 


24 X 6 


*320 


o 


2X2 


60 


0.32 


12 X 2 


4.7 




3X4 


55 


1.52 


15 X 3 


20 


* 2 


4X6 


55 


4.65 


20 X 4K 


53 




6X8 


50 


11.10 


30 X 6 


146 


.5 H 


8 X 10 


45 


21.00 


36 X 6 


292 



* Duplex double-acting pumps at the same speed give approximately 
twice these capacities and require twice the power and pulley width. 



94 



STEAM POWER PLANT AUXILIARIES 



[Div. 3 



Note. — Crank-And-Fly-Wheel Pump Sizes cannot be figured from 
the relative boiler pressure and working pressure as can direct-acting 

pump sizes (Sec. 50). Because of the 
cut-off at partial stroke of crank-and- 
fly-wheel pumps, the horsepower of 
the steam cylinders must be found 
and the capacity figured as for power 
pumps. 



Belt Idler 
Driving 
Motor- 




.'.o-o.'O.o; 

Fig. 117A. — Vaile-Kimes Single- 
Acting Deep- Well Pump Provided With 
Differential Piston For Securing Con- 
tinuous Discharge. (The plunger, C, on 
its up stroke discharges half of its dis- 
placement out the discharge, F, or into 
the air-chamber, H. The other half is 
drawn into differential cylinder, E, by 
the upward movement of D. On the 
down stroke, A closes and the water in 
E is discharged out F by D.) 



98. The Advantages Of The 
Electrically-Driven Pumping 
Unit are : (1) It may be located 
many miles from the source of 
power and still operate with very 
high efficiency. These values are 
typical: — Line efficiency, 90 per 
cent Motor, 85 per cent. Pump 
and gearing, 82 per cent. Over- 
all efficiency 63 per cent For 
steam or air-operated pumps 
which are installed a consider- 
able distance from the source of 
power, the over-all efficiency 
would probably be under 25 per 
cent. (2) Automatic control is 
effected readily with electricity. 
Electrically-driven pumps may 
be started and stopped by a 
float-operated switch which will 
maintain a required level in the 
supply tank. Electrically- 
operated pumps may readily be 
controlled from any reasonable 
distance. 



Note. — The choice of a method of 
driving a boiler feed pump is dis- 
cussed in Sees. 214 to 219. The prin- 
ciples outlined therein are of general 

application, and are useful in selecting driving means for a variety of 

purposes. 



Sec. 091 CRANK-ACTION PUMPS 95 

99. Simplex Double -Acting Pumps are manufactured for a 
great variety of purposes. Many non-corrosive oils, solutions 
and other liquids are handled in factories by such pumps. 
The sizes range ordinarily from around 2 in. bore and stroke to 
around 6 in. bore and stroke for general service. Small 
water-supply systems can often be served effectively by pumps 
of this simple type. Simplex pumps of small capacity have 
the advantages of lower first cost and greater ease of repair 
than more complicated pumps. In the larger capacities 
these advantages disappear. Simplex pumps are seldom de- 
signed single-acting because of the intermittent discharge due 
to such action. Single-acting deep-well pumps are an exception 
but the discharge is made regular in some such pumps by a 
differential cylinder, which is located near the discharge out- 
let and discharges half the water on the upstroke and half 
on the down-stroke (Fig. 117A). 

100. The Use Of Duplex Single -Acting Pumps is confined 
largely to a few special applications where it is necessary to 
reduce the first cost below that of a triplex pump. They are 
now made seldom, if ever. The intermittent discharge may 
be a decided disadvantage. For the average service, the 
duplex single-acting pump has no advantage over the standard 
simplex double-acting pump. 

101. Crank-And -Fly -Wheel Pumps range in size up to 
perhaps 10 ft. stroke by 4 ft. bore for municipal pumping 
service. The large pumps of this type are usually compound 
duplex or triple expansion triplex. Crank-and-fly- wheel pumps 
are, on account of their high economies, used occasionally 
for medium duty, although their first cost is greater than that 
of either the centrifugal or direct-acting pumps with which 
they are in competition. 

Note. — Centrifugal pumps driven by motors or steam turbines are 
superseding crank-and-fly-wheel pumps for large municipal pumping 
installations. The centrifugal unit usually deteriorates less in efficiency 
with constant use than does the reciprocating unit. Furthermore, the 
much smaller size and weight of the centrifugal unit for a given capacity 
make its installation less expensive. These features are conducive to 
lower annual costs. 



96 STEAM POWER PLANT AUXILIARIES [Div. 3 

102. Duplex Double-Acting Power Pumps are manufactured 
in sizes ranging from perhaps 3 in. bore, 4 in. stroke to 14 in. 
bore, 12 in. stroke for mine pumping, boiler feeding (in the 
smaller sizes), drainage and general water-supply purposes. 
The additional parts necessary for the two cylinders of these 
pumps are justified by the smaller size of the parts and the 
better characteristics of the duplex pump. The cranks of 
these pumps are usually set 90 deg. apart so as to give four 
maximum discharge peaks per revolution. 

103. Triplex Single -Acting Power Pumps are in competition 
with duplex double-acting power pumps for most classes of 
service and the choice of design varies with the manufacturer. 
The triplex single-acting is a more compact upright type of 
pump. The duplex double-acting type is more common in the 
horizontal design because of the extra length of guides neces- 
sary for double action. There is some advantage in the triplex 
single-acting construction for hydraulic press work because the 
strains are mere easily taken care of by the single-acting form 
of plunger and connecting rod. Triplex pumps are more 
commonly used than are duplex pumps. 

104. Triplex Double-Acting Pumps are used occasionally 
for certain special applications in large units for high-pressure 
pumping. For the average application they possess no advan- 
tage over single-acting triplex pumps. There are compara- 
tively few in use. 

Note. — Multi-stage centrifugal pumps are now used for many ser- 
vices where it was formerly considered that the head or pressure was too 
high for a centrifugal pump to work against. The efficiency of a cen- 
trifugal pump is usually somewhat less than that of a new crank-action 
pump. However, the centrifugal pump has advantages such as compact- 
ness, simplicity, low up-keep and long-continued efficiency that under 
many conditions offset this disadvantage. 

105. The One General Rule In Selecting A Pump is first 
to find which types of pumps will satisfy the capacity require- 
ments of the service being considered and be reliable under 
the conditions. In so doing, consider: (1) Liquid to be handled. 
(2) Attention required. (3) Characteristics. (4) Capacity, head 
and power. Then, the eligible types having been determined, 
select that type which will show the least annual cost or the 



Sec. 106] 



CRANK-ACTION PUMPS 



97 



least cost for pumping a certain quantity of water, on the 
basis of: (1) Interest on investment. (2) Depreciation. (3) 
Maintenance. (4) Power cost. Often this determination 
may be made most conveniently on the basis of pumping the 
quantity of liquid which the pump must handle in a year. 
106. Modern Pump Applications. — The words in the 
spaces (Fig. 118) refer only to crank-action pumps. It is 
understood that only one pump at a time is being considered. 
Greater capacities can be obtained, of course, by installing 
several pumps in parallel. Greater heads can sometimes be 




Fig. 118. — Modern Pump Applications. 

obtained by installing several pumps in series. The diagram 
should be studied in connection with Sees. 95 to 105. The 
letters S and D refer to single or double-action. The type 
names which are underscored, indicate the type ordinarily 
preferable for the stated conditions. 

107. To Compute The Horse Power Rating Which A Motor 
Should Have to Operate A Deep-Well Pump use the following 
formulas which were derived from data in the Goulds Mfg. 
Co. catalogue. 

When the pump operates single-acting (Fig. 101) or two- 
stroke (Figs. 103 and 104) : 

V gmlifimT 



(48) 



Pfc/ip — 



1,300 



(horse power) 



98 



STEAM POWER PLANT AUXILIARIES 



[Div. 3 



When the pump operates double-acting (Fig. 102) : 



(49) 



bhp 



V gm (LhmT + L/K) 



(horse power) 



2000 

Wherein : "P h h P = the required horse power. V gm = the quan- 
tity of water pumped in gallons per minute. Lh m r = the total 
measured head against which the pump works in feet. Lj = 
the length of the plunger rod in feet. K = sl constant taken 
from Table 108 by which the weight of the plunger rods and 
couplings is included in the computation. 

Note. — The quantity K is ignored in For. (48) because the weight of 
the single-acting rods which have to stand tension only is not great enough 
to enter into the calculation. The two-stroke pump plunger rods bal- 
ance as explained in Sec. 90. 

Example. — A two-stroke deep-well pump (Fig. 103) is required to 
deliver 100 gal. of water per minute against a total measured head of 
200 ft. What should be the horse-power rating of a motor which is to 
drive this pump? 

Solution.— By For. (48), V bhv = V gm L hm T /1300 = 100 X 200 -r- 
1300 = 15 h. p. approximately. 

Example. — A double-acting single-plunger pump is required to draw 
125 gal. per min. of water from a well 150 ft. deep and deliver it into a 
tank 100 ft. above ground. The bore of the pump cylinder is 5.75 inches. 
What should be the horse-power rating of a motor to drive this pump? 

Solution. — The total measured head = 150 + 100 = 250 ft. By 
Table 108, the value of K for a 5.75 inch pump = 0.56. By For. (49) 
Pwp = V gm (L hmT +L f K) /2000 = 125 X [250 + (150 X 0.56)] ^ 2000 = 
20.9 h.p., or 21 h.p. practically. 

108. Table Of Head-Pressure Equivalents K For. (49) Of 
Weight Of Deep-Well Pump Plunger Rods. 



Dia. pump 


K or head per 


Dia. pump 


K or head per 


cyl. inches 


ft. of rod 


cyl. inches 


ft. of rod 


2.25 


0.96 


4.75 


0.59 


2.75 


0.69 


5.75 


0.56 


3.25 


0.72 


6.50 


0.46 


3.75 


0.73 


7.50 


0.36 


4.25 


0.60 


8.50 


0.40 



109. Leather Cup-Washers For Deep-Well Pump-Plungers 

should be of the best quality of oak tanned leather. Soft 
spongy leather is utterly unsuited for this service. 



Sec. 110] 



CRANK- ACTION PUMPS 



99 



Note. — Leather packing should be thoroughly greased with pure tal- 
low. The tallow should be worked into the leather with the fingers be- 
fore the cup is put into place. Satisfactory lubrication may also be 
secured by soaking the cups in neatsfoot, sperm, or castor oil for an hour 
before putting them into place. In no case should mineral oil be used. 
Treatment with ordinary machine oil, 
which contains a mineral ingredient, 
tends to rot the leather and render it 
pulpy. 

Note. — To Make A Set Of Cup- 
Washers For A Pump Plunger, 
proceed as shown in Fig. 119. The 
cast-iron mould, M, should be made 
with d\ equal to the diameter of the 
pump cylinder and S }>i 2 in. greater 
than the thickness of the leather, 
(Table 110). The radius of the 
mould at R should be about one 
third the height of the washer. A 

disk of leather, the proper diameter and thickness, is soaked in water until 
soft. Then it is drawn down slowly into shape by means of the 
bolt. The protruding edge is then trimmed off flush with the matrix. 
After ten hours or more, the leather is removed and well greased with 
tallow. 




Drainage Duct- : Clamping Sc't- 



Fig. 119. 



-Mold For Forming 
Leathers. 



Cup- 



110. Table Of Dimensions Of Cup-Washers For Pump 
Plungers. 



Diameter of pump 
cylinder in inches 



Thickness of 
leather in inches 



D 

Depth of cup 
in inches 



2 


He 


5 A 


3 


He 


H 


4 


H 


l 


5 


H 


m 


6 


H 


IK 



QUESTIONS ON DIVISION 3 

1. What are the two principal classes of crank-action pumps? Define each. 

2. Why may steam be used expansively in crank-and-fly-wheel pumps and not in 
direct-acting pumps? 

3. Give values for the steam consumption of high-duty crank-and-fly-wheel pumps, 
run condensing. Non-condensing. 

4. What are the disadvantages of crank-and-fly-wheel pumps, as compared to direct- 
acting steam pumps? 



100 STEAM POWER PLANT AUXILIARIES [Div. 3 

5. What two kinds of deep well pumps force water up the drop pipes in a fairly- 
continuous stream? What kind does not? Can this last kind be made to give fairly 
continuous discharge? How? 

6. Explain, with a sketch, the operation of a double-acting single-plunger deep-well 
pump. Of a two-stroke pump. 

7. What do the graphs of Figs. 112, 113 and 115 represent? What do they show 
about the action of various kinds of pumps? 

8. Under what condition can alcohol and gasoline be pumped satisfactorily? 

9. Give several advantages of electric dr've for a remotely located pump. 

10. What type of pump is superseding the large crank-and-fly- wheel pump? Why? 

11. What is the advantage of the simplex double-acting pump for small capacity 
requirements? 

12. Name two widely-used types of crank-action power pump other than the simplex 
double-acting type. Which of the two is most commonly used? 

13. Outline a method of arriving at a proper choice of power pump. 

14. What is a cup washer for? Explain by a sketch how to make one. How should 
it be lubricated? 

PROBLEMS ON DIVISION 3 

1. Compute the proper horsepower rating for a motor which is to drive a single- 
acting deep-well pump delivering 150 gal. per min. against a total measured head of 
225 ft. 

2. Compute the proper horsepower rating for a motor which is to drive a double- 
acting deep-well pump having a displacement of 0.9 gal. per rev. at 30 r.p.m. The 
well rod is 175 ft. long and the pump discharges 50 ft. above the base of the generating 
gear. Cylinder diameter is 2% in. 



DIVISION 4 



CENTRIFUGAL AND ROTARY PUMPS 



(Discharge Out/ef 
' ^ Cas, 

^2/ 



111. The Development Of The Centrifugal Pump started 
with its invention in about 1680. The first centrifugal pump 
built in America (Fig. 120) was called the Massachusetts 
pump. This was a crude affair of low efficiency. Only 
during the last 20 years has much improvement been made 
over the Massachusetts pump. This seemingly slow develop- 
ment has been due to the fact 
that the centrifugal pump is 
inherently a relatively - high - 
speed machine. Formerly, there 
was no motive power well adapted 
to drive it. The introduction of 
the electric motor and the steam 
turbine, which are inherently 
high-speed machines, led to 
further development. Hence the 
demand for centrifugal pumps 
is now great and is steadily 
increasing. 




Fig. 120. — The Massachusetts 
Pump. 



Note. — A large portion of the material contained in this Div. is based 
on that from publications of The Goulds Manufacturing Co., to whom 
credit is hereby given. 

112. A Centrifugal Pump is a pump that, as will be ex- 
plained later, depends upon centrifugal force or the variation 
of pressure due to rotation for its action. When any body 
is constrained to move in a curved path, there is a force which 
tends to impel the body outward from the center. This 
force is called centrifugal force. 

113. The Theory Of The Centrifugal Pump may be illus- 
trated (Figs. 121 and 122) by the phenomenon of a bucket of 
water which is whirled around the head in a circular path. 
If the bucket of water is whirled at a sufficiently-high speed, 

101 



102 



STEAM POWER PLANT AUXILIARIES 



[Div. 4 



none of the water will spill, even when the bucket is in the 
position shown in Fig. 121. The force which holds the water 
against the bottom of the bucket is centrifugal force. Now if 



^'Centrifugal Forc£\ 
Presents Wafer 
From 



<rs water \ . 
5pf///ng ~*i 




DirectionOf Centrifugal Force Is Radially 
■Outward From Center 



■Rapidly 
Whirling 
Bucket 




Circular Path Of 
Bucket-^ 



Fig. 121. — Illustrating Centrifugal Force. Fig. 122. — Centrifugal Force Holds The 

Water Against Bottom Of Bucket. 



— Centrifugal """^ 
Force Causes *\. 
Water To Pass \ 
Through Hole In ia 

Bottom Of Bucket \ 
And Upward \ 
'.Into The Air 




Fig. 123. — Illustrating The Principle Of 
The Centrifugal Pump. 



Direction Of Centrifugal Force Is Rad- 
ially Outward From Center 




\ Circular Path Of / 

\ Bucket- / 



■ -« —* 

Fig. 124. — Showing That Centrifugal 
Force Causes The Water To Flow Out- 
ward Through The Hole In The Bucket. 



a hole is cut in the bottom of the bucket, the water will be 
forced out through the hole (Figs. 123 and 124) and will be 
thrown upward into the air. 



Sec. 114] CENTRIFUGAL AND ROTARY PUMPS 



103 



Explanation. — Suppose that the boy's arm is a suction pipe and that 
his bocfy is a reservoir containing water (Fig. 125). The centrifugal 
force of rotation throws the water from the bucket. This tends to pro- 
duce a vacuum within the bucket and 

suction pipe. If the surface of the .Water Thrown' "? M ^/cJT/x 

. f? . . , xl { From Buckef'.-Certrifugraf Force ^ N 

water in the reservoir is open to the y__ /f x \ 

atmosphere, the water will be forced to 1 



Arm).. 




Atmospheric 
Pressure- ■• 



i±± 



-Pivot \ 

I 
Path Of Potato 
ing Bucket-^ / 
/ 
/ 
/ 
/ 



Reservoir Contain- 
ing Waterboys 
y Body^ 



rise in the suction pipe by the atmos- 
pheric pressure and will be pushed by 
the centrifugal force out through the 
hole in the bucket so long as the end 
of the pipe (Fig. 125) is submerged in 
the water and the bucket is rotated at 
a sufficiently-high speed. Roughly, 
this illustrates the theory of the cen- 
trifugal pump. 



114. The Commercial Cen- 
trifugal Pump (Fig. 126) is 
merely a modification of the ap- 
paratus shown in Fig. 125. The 
impeller, rotating within the cas- 
ing, C, corresponds to the rotating 
bucket. Water enters the impeller through an inlet hole 
around its center, 0. The rotation of the impeller imparts cen- 
trifugal force to each particle of water, which causes the 



\j,,UJ>//>//*/>/SJ. 



Fig. 125. — Illustrating The Principle 
Of The Centrifugal Pump. 




Suction... 
If ■ 



t II \\ 

■Yoiufe 
IL-Trotmvirse section 



I- Section aa 
Fig. 126. — Single-Stage, Single-Suction Volute Centrifugal Pump 



water to be thrown outward. Thereby pressure is created 
back of each particle of water and the water is discharged 
from the impeller into the case, C. The contour of the im- 



104 



STEAM POWER PLANT AUXILIARIES 



[Div. 4 



peller blades is so designed that the water enters the 
blades, passes through them and is discharged with a minimum 
of friction. 

Explanation. — The water upon entering the pump at 0, (Fig. 127) 
is caught between the vanes of the impeller which are rotating. This 
rapid rotation of the water sets up a centrifugal force, F, and forces the 
water outward against the pump casing, C, just as the boy swinging the 
bucket over his head (Fig. 121) created a centrifugal force which pressed 
the water against the bottom of the bucket. The pressure which is 

Discharge Pipe 1 Sq. In. In Cross Szcffon 

Wafer Enters Pump Through This 

..-Suction Pipe 



Direction Of"-. 
Rotation • 




Fig. 127. — Centrifugal Force Created By Rotation Of Impeller Vanes. 



thus set up may be imagined to be transmitted by the water, from 
particle to particle, entirely around the inner periphery of the casing to 
the discharge nozzle. The water is thus caused to rise in the dis- 
charge pipe P, just as the water was forced out through the hole in the 
bottom of the rotating bucket. The water will rise in the pipe until the 
pressure due to the water column in P just balances the centrifugal force 
F. Suppose the speed of rotation of the impeller /, (Fig. 127) is such that 
a centrifugal force of 43.4 lb. per sq. in. is produced on the casing. Sup- 
pose the nozzle and discharge pipe, P, have a cross-sectional area of 
1 sq. in. Water will then rise in the discharge pipe until the weight of 
the water column is 43.4 lb. The height of a water column 1 sq. in. in 
cross section having a weight of 43.4 lb. is (Sec. 5) 100 ft. It will be 



Sec. 115] 



CENTRIFUGAL AND ROTARY PUMPS 



105 



shown later that the impeller velocity which is required to lift water 
vertically 100 ft. is the same as that velocity which the water would have 
after freely falling through a distance of 100 ft. 

Note. — There Are Other Factors Which Have Considerable 
Effect upon the efficient operation of centrifugal pumps, such as elimi- 
nation of eddy currents, efficient transformation of kinetic energy to 
pressure without shock, etc. These are principles of design and are 
not within the scope of this book. 

115. A Freely Falling Body Will, If It Falls Through A 
Certain Height, Have A Certain Velocity, or speed, at the end 



^Vessel Of Wafer 




Wmmz 



Fig. 128.— Vessel Fall- 
ing From Top Of A 100-Ft. 
Building. 



of its fall. Suppose there is a body, say 
a bucket of water, on the top of a building 
(Fig. 128) which is 100 ft. high. If the 
bucket is pushed off and allowed to fall, 
it will fall with a continuously increasing 
speed until it strikes the earth. If it is 
now impelled upward with an initial 
velocity equal to the velocity which it 
had when it struck the earth, it will 
rise just to the height from which it fell. 
The velocity which a body will acquire 
in falling through a given distance, or 
the velocity which must be imparted to 
a body to cause it to rise to a given 
height may be computed by the follow- 
ing, which is, if the frictional resistance of the air be disre- 
garded, true for any body whatsoever: 

(50) v = V2gL f (ft. per sec.) 
or 

(51) v m = 481 Vi/ (ft. permin.) 

Wherein: v = velocity in feet per second. v m = velocity in 
feet per minute, g = acceleration due to gravity = 32.2 ft. 
per sec. per sec. L } - = distance in feet, through which body falls, 
or the height to which it will rise if impelled upward with an 
initial velocity of v or v m . 

Example. — A vessel of water (Fig. 128) is dropped from a point 100 
ft. above the earth. With what velocity will it strike the ground? 
Solution. — By For. (51), the velocity = v m ='481VL/ = 481\/l00 = 
481 X 10 = 4,810 ft. per min. 



106 



STEAM POWER PLANT AUXILIARIES 



[Div. 4 



Example. — What is the initial velocity which must be imparted to 
the vessel of water to cause it to rise 100 ft. in a vertical direction? 
Solution. — By For. (51), the velocity = v m = 48 ly/Lf = 481\/l00 = 
481 X 10 = 4,180/i. per min. 

116. The Theoretical Speed In R.P.M. At Which A Cen- 
trifugal Pump Impeller Must Run To Pump Water To A Cer- 
tain Height may be determined by The Law Of Freely Falling 

Bodies. As was shown in the 
preceding Sec, the water, to be 
thrown to a certain height, must 
have the same velocity when it 
leaves the impeller as it would 
have if it fell from the same 
height. This may be stated: 
The speed in feet per minute of 
a point on the periphery of the 
impeller should be equal to the 
velocity which the water would 
acquire in falling from the same 
height as the total head pumped 
against. 

Note. — The Total Head Pumped 
Against is the sum of all friction, 
velocity, and static heads, which occur 
between the suction-pipe intake and 
the delivery-pipe outlet. See Sec. 12. 
Example. — At what speed, in r.p.m. 
must a 12-in. diameter impeller of a 
centrifugal pump (Fig. 129) be driven 
to deliver water against a total head 
of 121 ft.? Solution— -By For. (51), 
velocity = v m = 481-\/£/ = 481^/121 = 
481 X 11 = 5,291 ft. per min., which is the required peripheral velocity 
of the impeller. Circumference of impeller = ird = 3.1416 X 1 = 3.14 /£., 
which is the distance a point on the periphery of the impeller will travel 
during 1 revolution. Now, 3.14 X r.p.m. = peripheral velocity of the 
impeller = 5,291. Or, r.p.m. = 5,291 -f- 3.14 = 1,685 r.p.m. 

Note. — Due to certain losses which cannot be eliminated, the actual 
speed of the impeller must be somewhat greater than the theoretical 
speed to produce a given head. 

Note. — "Head" May Be Reduced To Equivalent Pounds Per 
Square Inch Unit Pressure as explained in Sec. 4. Also see the 
author's Practical Heat for definition and explanation of unit pressure. 




Fig. 129. — A 12-In. Diameter Im 
peller When Driven At 1685 R.P.M 
Will, Theoretically, Produce A 121-Ft 
Head. 



Sec. 117] CENTRIFUGAL AND ROTARY PUMPS 107 

117. The Quantity Of Water Which A Pump Will Deliver 

when being driven at a given speed will depend upon: (1) The 
size of the discharge outlet. (2) The size of the suction inlet. (3) 
The size of the casing. (4) The width of the impeller vanes. In 
good design the allowable velocity of the water at the discharge 
outlet is about 10 ft. per sec. However, this velocity may 
vary from 5 to 15 ft. per sec. 

Note. — It Is Customary, In Ordinary Parlance, To Speak Of A 
Centrifugal Pump As A "4-in. pump," sl u 6-in. pump," Etc. This 
means that the inside diameter of the discharge nozzle, N, Fig. 126, is 
4 in. or 6 in. However, the discharge-nozzle diameter is not to be taken 
as accurately defining the capacity of a pump. But if it is remembered 
that the nozzle-velocity in most centrifugal pumps is about 10 ft. per 
sec, the discharge-nozzle diameter does provide some idea as to the 
capacity of the pump in gallons per minute. An approximate rule is: 
The number of gallons discharged per minute is approximately equal to the 
square of the discharge-nozzle diameter, in inches, multiplied by 25. 

118. The Quantity Of Water Delivered By A Centrifugal 
Pump Through A Frictionless Pipe Will Vary In Direct Pro- 
portion To The Speed Of The Impeller, If The Diameter of the 

impeller remains unchanged, and if the friction of the water in 
the pump is neglected. This may be formulated as follows: 

(52) V gm2 = — ^ — °— (gal. per min.) 

Wherein: V gm2 = quantity of water, in gallons per minute, 
delivered by the pump when running at N 2 r.p.m. V gm i = 
quantity of water delivered by the pump when running at 
Ni r.p.m. 

Example. — A certain centrifugal pump running at 1,600 r.p.m. de- 
livers 1,000 gal. per min. through a frictionless pipe line. How many 
gallons will be delivered per minute by the same pump through the same 
pipe if the speed is changed to 1,200 r.p.m. Solution. — By For. (52), 
the quantity which will be delivered at the changed speed = V gm 2 = (Nz X 
V am i) -5- Ni = (1,200 X 1,000) -J- 1,600 = 750 gal. per min. 

Note. — Since All Actual Pipe Lines Offer Frictional Resist- 
ance To Water Flow In Them, The Above Formula Cannot Be 
Used In Practice. The actual quantity of water delivered by a pump 
through a pipe line may be either greater or less than the value obtained 
by applying the above formula. The only practical method of deter- 
mining the delivery of an actual pump at different speeds is by test, as 
explained in Sec. 138. 



108 STEAM POWER PLANT AUXILIARIES [Div. 4 

119. The Pressure Head Which Will Be Produced By A 
Centrifugal Pump Will Vary As The Square Of The Speed Of 
The Impeller, if the diameter of the impeller remains constant 
and there is no water-friction loss within the pump. This may 
be expressed as a formula by: 

(53) L hT2 = (^) L m (feet) 

Wherein: L hT 2 = head, in feet, produced by the pump when 
running at iV 2 r.p.m. L hT i = head, in feet, produced by the 
pump when running at Ni r.p.m. 

Example. — A pump which has no water-friction loss is running at 
1,600 r.p.m. produces a total head of 80 ft. What head will be produced 
by the same pump if the speed of the impeller is changed to 1,000 r.p.m.? 
Solution. — By For. (53), the head produced at the new speed = LhTi = 
(N 2 + NO 2 X L h Ti = (1,200 -r 1,600) 2 X 80 = % 6 X 80 = 45 ft 

120. The Power Required To Drive A Centrifugal Pump 
Will Vary As The Cube Of The Speed Of The Impeller, if 

the diameter of the impeller remains unchanged, and if no 
power is lost through pump by mechanical and water friction. 
This rule may be written: 

(54) V bhp2 = (-^- 2 j Pbh P i (horse power) 

Wherein : P & ^ p2 = horse power required to drive the pump at a 
speed of N 2 r.p.m. P^pi = horse power required to drive the 
pump at a speed of Ni r.p.m. 

Example. — 32 h.p. are required to pump a given quantity of water 
against a certain head when the frictionless pump is running at 1,600 
r.p.m. What would be the horse power required to drive the same pump 
at 1,200 r.p.m.? Solution. — By For. (54), the power required at the 
new speed = P bhp2 = (N 2 -*- Ni) 3 X P bhP i = (1,200 -i- 1,600) 3 X 32 = 
27/64) X 32 = 13.5 h.p. 

121. The Velocity Of A Point On The Periphery Of The 
Impeller Is Directly Proportional To The r.p.m. Of The 
Impeller, or expressed as a formula : 

(55) v m = N * 2 * d = 0.261,8 Nd (ft. permin.) 



Sec. 122] CENTRIFUGAL AND ROTARY PUMPS 109 

Wherein : v m = velocity, in feet per minute, of a point on the 
periphery of the impeller. JV = speed, in r.p.m., of the 
impeller, d = diameter of the impeller in inches. 

Note. — By transposing For. (55) and substituting in Fors. (52), (53), 
and (54), there results: 
From For. (52) 



(56) V gm2 = ■ 




(gal. per min.) 


From For. (53) 






(57) Ut-i = 


(l!) 2jLm 


(feet) 


And from For. (54) 






(58) Vbh P 2 = 1 


(I) 3 Pbhpl 


(horse power) 



Wherein: di and d 2 = the old and new diameters of the impeller, in 
inches, respectively. From Fors. (56), (57), and (58), it is evident, that 
if the speed in r.p.m. of a centrifugal-pump impeller remains constant, 
and if there is no friction, the following will be true: (A) From For. (56), 
the quantity of water delivered mill vary as the diameter of the impeller. (B) 
From For. (57), the head produced will vary as the square of the impeller 
diameter. (C) From For. (58), the power required for driving will vary as 
the cube of the impeller diameter. 

122. Centrifugal Pumps May Be Classified According To 
Several Different Features, the most important of which are : 
(1) Volute or turbine. (2) The number of stages. (3) Single 
suction or double suction. (4) Open impeller or enclosed 
impeller. (5) Horizontal or vertical. Each of these different 
features will be discussed in succeeding Sees. 

123. The Two General Classifications Of Centrifugal 
Pumps Are: (1) Turbine Pumps. (2) Volute Pumps. The 
turbine pump (Fig. 130) is one wherein the impeller is sur- 
rounded by a diffusor containing diffusion vanes which direct 
the water flow from the impeller. The relative position of 
the diffusor, D, and the diffusion vanes, V, (also called guide 
vanes) is shown in Fig. 131. These vanes are so shaped that 
gradually enlarging passages are provided for the water. 
In flowing through these guide-vane passages, the velocity 
which is imparted to the water by the centrifugal force 
(Sec. 114) is converted into pressure. The casing which sur- 
rounds the diffusion ring, may be circular and concentric 



110 



STEAM POWER PLANT AUXILIARIES [Div. 4 




Sec. 124] CENTRIFUGAL AND ROTARY PUMPS 



111 



(Fig. 131) with the impeller, or is sometimes of a spiral form. 
The volute pump (Fig. 126) 
is one which has no guide 
vanes, but instead, has a 
spiral-shaped casing. This 
spiral casing F,(Fig. 126) 
is also called the volute. 
In the volute pump, this 
spiral casing replaces the 
guide vanes of the turbine 
pump. The volute, or 
spiral casing, is so de- 
signed that it so guides the 
water from the impeller to 
the discharge pipe that the 
velocity is gradually con- 
verted into pressure. Vol- 
ute pumps ordinarily have 
but a single impeller. 
Where a closed-type im- 
peller is used, a double- 
inlet is employed, thereby 
eliminating end thrust. 

124. The Applications Of 
The Volute Pumps And Of 
The Turbine Pumps over- 
lap. In general, however, 
for low heads, (under about 
70 or 80 ft.) the volute 
pump should be chosen. 
For higher heads the tur- 
bine (multi-stage, Sec. 
126) pump will give better 
service. The volute pump 
may be considered superior 
to the turbine pump from 
the standpoint of size, sim- 
plicity, and cheapness. 




112 



STEAM POWER PLANT AUXILIARIES 



[Dw. 4 



Note. — There Is Much Controversy Concerning The Compara- 
tive Efficiency Of The Two Types Of Pumps. More rapid progress 
has probably been made in the design of the turbine pump than in that 
of the volute pump. This is attributed to the fact that the guide-vane 
design of the turbine pump is more amenable to mathematical analysis 
than is the spiral casing of the volute pump. It has been predicted, 
that volute passages will eventually be designed whereby it will be possi- 
ble to effectively pump against the same heads with the volute pump as 
with the turbine pump. Since the volute pump is the cheaper and 
simpler it may therefore find a wider application in the future. 

125. Water May Be Raised As High As Desired by arrang- 
ing a sufficient number of independent pumps (Fig. 132) so 
that the discharge of one of the pumps is piped to the suction 

of the next. It is desired to 
pump the water (Fig. 132) to 
a total height of 200 ft. 
Pump A takes water from 
reservoir D and delivers it 
to reservoir E. Pump B 
takes water from reservoir 
E and delivers it to reservoir 
F. This is, however, an un- 
economical method of pump- 
ing water against a high 
head. The usual method 
which is used. in practice is 
described in the following 

(jC ; TTy^ P . 1 SeC ' 

CSS/ 2 126. The Multi-Stage 

Centrifugal Pump (Figs. 130 
and 133) is really two or 
more distinct pumps con- 
nected in series. Such a 
pump has two or more im- 
pellers through which the water passes successively. The 
impellers are mounted on the same shaft and contained within 
the same casing. That is, the water is discharged from the 
first-stage impeller, I, (Fig. 133) through the return chamber, 
R\ to the suction side of the second-stage impeller, II, etc., 
throughout each stage of the pump. Multi-stage pumps are 




Fig. 132. — Showing How Water May Be 
Pumped To A Great Height By Separate 
Steps Or Stages. 



Sec. 125] CENTRIFUGAL AND ROTARY PUMPS 



113 




114 STEAM POWER PLANT AUXILIARIES [Dw. 4 

used to pump against high heads. They may be either of 
the volute or of the turbine type. 

Explanation. — The two-stage pump, (Fig. 130) may be considered 
merely as a more compact arrangement of the two pumps in Fig. 132. 
Suppose the water is taken into the first-stage suction, Si (Fig. 130) and 
is discharged to the second-stage suction, S 2 , through the return chamber, 
R, at a pressure equivalent to a 100-ft. head. The water is then received 
by the second-stage impeller under a 100-ft. head. In passing through 
the second-stage impeller, the water is given an additional 100-ft. pres- 
sure head. Thus as the water passes from Si to N (Fig. 130) the same 
result is obtained as by the two pumps in Fig. 132. Multi-stage pumps 
are usually designed to produce from about a 100- to a 150-ft. head per 
stage. The superiority due to compactness, simplicity, and economy of 
the multi-stage pump of Fig. 130 over the two-pump arrangement of 
Fig. 132 is obvious. 

127. "Single Suction" And "Double Suction" are also 
classifications of centrifugal pumps. A single-suction (also 
called side-suction) pump (Fig. 133 is one in which the water 
enters the impeller from one side only. A double-suction 
pump (Fig. 130) is one in which the water enters the impeller 
from both sides. A double-suction pump will, with same im- 
peller diameter, have a larger discharge than a single-suction 
pump. The double-suction pump may have two separate 
suction pipes, or the water may be divided after it enters the 
casing. A single-suction pump which takes water, either by 
suction or under a positive head, will have a side-thrust. Side- 
thrust is caused by the pressure on one side of the impeller 
being greater than the pressure on the other side. This side- 
thrust is transmitted to the shaft, and will, unless some method 
of balancing is provided, cause excessive friction and wear in 
the thrust bearing. 

128. The Forces Which Tend To Unbalance The Impeller 
may be understood from a consideration of Fig. 134. The 
water, which enters the impeller eye at A, has its direction of 
flow parallel to the axis of the shaft. When the water im- 
pinges on the impeller at B, its direction of flow is changed, 
as shown by the arrows. This change of direction results 
in the exertion of a force against the impeller which tends to 
move it to the right. Since the pressure in pounds per square 
inch in r is almost equal to the pressure in pounds per square 



Sec. 129] CENTRIFUGAL AND ROTARY PUMPS 



115 



Impe/fa 



inch at the periphery of the impeller, the water in r will exert 
a force on the impeller, the direction of which will be to the 
left. Due to the same cause, a pressure will exist in t, which 
will exert a force to the right on the impeller. However, the 
leakage of water through s will result in the pressure in pounds 
per square inch in t being somewhat less than that in r. Also, 
the area of the impeller web over which the force in r acts 
is greater than that over which 
the force in t acts. Therefore, 
since the pressure in pounds per 
square inch in t is less than that 
in r, and since the area of r is 
greater than that of t, the 
combined - transmitted - pressure 
force will act to the left on the 
impeller. As all of these forces 
may vary 'from one instant to 
the next, the direction of the 
resultant may shift from right 
to left. It cannot, therefore, 
be predetermined just how great 
or in which direction the result- 
ing force will be. To minimize 
the total resultant unbalance 
the devices which will be de- 
scribed are employed. 

129. There Are Various Methods Of Balancing Single- 
Suction Impellers Against End-Thrust, the most common of 
which are: (1) The Jaeger method. (2) By means of an auto- 
matic hydraulic balancing piston. Each will be described: 

130. The Jaeger System Of Balancing Single -Suction 
Impellers (Fig. 135) automatically minimizes the longitudinal 
unbalance but it requires, in addition, mechanical thrust 
bearings. The impeller is equipped in front and rear, with 
wearing rings (R, Fig. 135). The diameter of the front and 
back rings is the same, so that the area of the surface a is 
equal to that of surface b. Since leakage through the rings 
will be practically the same in both the front and the back 
sides, the pressure on a will be equal and opposite to that on 




Fig. 134. — Unbalanced Impeller. 



116 



STEAM POWER PLANT AUXILIARIES 



[Div. 4 



b. The leakage water which flows across the front sealing 
surface enters the suction opening of the impeller. To prevent 
the leakage water which flows across the back sealing surface 
from collecting in the annular ring and building up pressure, 
the holes, H (Figs. 131 and 135), permit 
this leakage water to pass into the im- 
peller. Leakage through the wearing 
rings may be minimized by forming a laby- 
rinth pathway (Fig. 136) for the water. 
The mechanical thrust bearings which are 
necessary to resist the force due to change 
of direction (Sec. 128) are usually of the 
ball type, or (Fig. 130) of the multi-collar 
type. 

131. The Automatic Hydraulic Balanc- 
ing Piston (Fig. 133) whereby all of the 
impellers (multistage pump) are balanced 
by a single balancing piston is shown in 
Fig. 137. This balancing chamber is at 
the right-hand end of the last stage. The 
last-stage impeller is provided with wearing 
rings. Water leaks through between the surfaces of these 
wearing rings to the balancing chamber. If the shaft, and 
the movable part, M, (Fig. 137) moves to the right, the pas- 
sageway between the wearing rings is increased. This permits 
the water to pass more freely through R into the balancing 




- Showing 
Jaeger Method Of Im- 
peller Balancing. 



'-Casing—. 




f-Flat Type 

Fig. 136. — Various Types Of Wearing Rings 

C = Casing.) 



Impeller-' 
II -Off set Type HT-Labyrinth Type 




(7 = Impeller, W = Wearing Ring, 



chamber C. This same movement to the right tends to close 
the escape-passageway, E, which prevents the water from 
escaping through the pipe, P. Thus, the pressure in the 
balancing chamber builds up and acts against the balancing 



Sec. 132] CENTRIFUGAL AND ROTARY PUMPS 



117 



disk, (or piston) D, which is fixed to the shaft. This moves 
the shaft to the left until R is closed and E is open and 
equilibrium is established. 

Note. — Balancing Of Double-Suction Pumps is taken care of, 
theoretically, in the design of the pump. The liquid is supposed to 
enter in equal volumes from both sides. Since the inlet openings are 
also supposed to be equal, the vacuum or pressure on one side of the 
impeller is always equal and opposite to that on the other side. There- 
fore, no end-thrust is exerted. The impeller is also equipped with front 
and back wearing rings of equal diameter (Sec. 130) so that there is no 
end-thrust on the impeller on the outside of the wearing rings. Actually, 



Last- 
Stagz 



Pipe To F/rsf-Sfage,- Suction ChotmbZK 




Fig. 137.- 



-Piston, Or Automatic, Balancing System For Centrifugal Pumps. 
Laval Steam Turbine Co.) 



(De 



however, the inlet-openings are never exactly equal. The wearing rings 
are likely to wear unevenly. One or both of these causes will set up an 
unbalanced end-thrust on the inside of the wearing rings, making it 
necessary to equip a pump of this type with a mechanical thrust bearing. 
Note. — Due To The Small Bearing Surface Of The Open-Type 
Impeller (Fig. 127) very little end-thrust is developed. Hence, mechan- 
ical thrust bearings will ordinarily assume the end-thrust which is devel- 
oped in a pump of this type. 

132. The Open Impeller is shown in Figs. 138 and 138A. 
Pumps equipped with an impeller of this type are sometimes 
called fan pumps. The action is similar to that of a paddle- 
wheel revolving in a circular casing. All of the early centri- 
fugal pumps were of this type. It has poor water-guidance 
and flow-lines. This results in excessive wasteful churning 
and eddying of the water. Also, a great amount of water 



118 



STEAM POWER PLANT AUXILIARIES 



Div. 4 



escapes between the blades of the impeller and the casing walls. 
This is similar to the slip (Sec. 22) in reciprocating pumps. 



A«-i 




Blade- 



Bos s- 




E-Sec t ion A-A 

Fig. 138. — Open Type Of Impeller. (Pumps for use as power-plant auxiliaries are 
seldom equipped with open impellers.) 

Due to the above mentioned causes, the efficiency of the pump 
which is equipped with an open impeller is comparatively low. 
It is relatively cheap in price. For certain classes of work 




Fig. 138A. — Perspective View Of An Open-Type Centrifugal-Pump Impeller. 

such as pumping mash and thick liquids, it is the only type of 
centrifugal pump that will give satisfaction. Its use is not 
to be recommended as a power-plant auxiliary. 



Sec. 133] CENTRIFUGAL AND ROTARY PUMPS 



119 



133. The Enclosed Impeller (Fig. 139) is a development of 
the open impeller. If a disk or plate were secured to each 
side of an open impeller, a closed impeller would result. The 
enclosing walls or covers are, in practice, cast solid with the 
impeller vanes. These enclosing walls prevent the water 
from escaping past the impeller blades. Also a relatively 
close-running joint can be made between the impeller and the 
casing. This reduces the slip to less than that which occurs 
with the open impeller. The efficiency of the pump is mate- 



Vo/nes-. 




I- Side view I- Sectional Elevation 

Fig. 139. — Closed-Type Impeller. 

rially increased by these two devices. The running joint (Fig. 
135) is usually known as the sealing surface. The running 
joint is formed by the wearing rings. 

Note. — The Ideal Condition Would Be To Have A Tight Fit 
Between The Sealing Surfaces. This is, however, impossible of 
attainment. In practise, a diametral clearance of from 0.012 to 0.018 
in., is allowed between the wearing rings. Small particles of grit in the 
water will cause the rings to wear, thus enlarging the clearance and in- 
creasing the leakage. The increased leakage will lower the efficiency. 
This necessitates renewing of the wearing rings. 

134. The Maximum Heads Against Which Impellers Of 
The Different Types Are Designed To Operate are approxi- 
mately as follows: (1) Single-suction, open impeller, 100 ft. 
(2) Double-suction, open impeller, 100 ft. (3) Single-suction s 



120 



STEAM POWER PLANT AUXILIARIES 



[Div. 4 



enclosed impeller, 100 ft. 
150 ft. 



Dr/vingf 
Motor- 




^^ 




(4) Double-suction, enclosed impeller, 



Note. — The Single Im- 
peller Pump (R. A. Fiske) 
may be efficiently used for 
heads up to and including 150 
ft. or higher, with efficiencies of 
from 50 to 80 per cent. For 
pressures above 50 lb. per sq. 
in., two or more runners or 
stages may be used, each stage 
adding approximately 50 lb. per 
sq. in. to the total pressure 
available from the pump. 



Gu/c/e 
Beoirincf 



r=rfj2^£ Wi 



zrapm 

him 

lis 

.^•'■0. fc|p° 

#13? 






135. A Vertical -Shaft 
Centrifugal Pump (Fig. 
140) may be used where 
conditions are such that a 
horizontal-shaft pump can- 
not be placed within suc- 
tion distance of the 
supply-water level. This 
condition is frequently en- 
countered in deep wells, 
sewage service, sumps, 
and along rivers where the 
difference in water level 
between high and low 
water will amount to 20 
or 30 ft. A vertical cen- 
trifugal pump may be 
operated completely sub- 
*o^|If,l « merged in the water (Sec. 
136). It is, however, ad- 
visable, where conditions 
permit, to locate the pump 
in a dry-pit. This makes 
it more readily accessible 
than when submerged. 
Consequently the pump will be given better attention. For 



"W 






m 






Fig. 140. 



-A Vertical Centrifugal Pump Of 
The Submerged Type. 



Sec. 136] CENTRIFUGAL AND ROTARY PUMPS 



121 



reasons, which will be stated in the following Sees., a vertical- 
shaft centrifugal pump should not be selected where it is 
feasible to use one of the horizontal-shaft type. 

136. The Bearings In A Vertical Pump are very likely to be 
a source of constant trouble. These bearings may be divided 



Oil- 
Drain Plug--'* 




Adjustment Bolt 




Fig. 141. — Sectional View Of Hanger- Type 
Thrust Bearing For Vertical Centrifugal 
Pumps. (Worthington Pump And Machinery 
Corp.) 



Fig. 142. — Showing Rotating 
Parts And Thrust Bearing Of A 
Vertical Centrifugal Pump. (The 
Goulds Mfg. Co.) 



into two classes: (1) The pump bearings proper. (2) The line- 
shaft bearings. The pump bearings, if the pump is submerged, 
usually depend upon the water for lubrication. This results 
in extremely rapid wear. The line-shaft bearings consist of 
the thrust bearings (Figs. 140, 141 and 142), and, if the line 
shaft is long, the guide bearings (Fig. 143). The thrust bear- 
ing must carry the weight of the rotating parts and, in some 
instances, the weight of the pump. It has been found difficult 



122 



STEAM POWER PLANT AUXILIARIES 



[Div. 4 



Vertical 
Shaft-' 



BctbbitK 



Oil Forced Up 
Through Plan 
By Centrifugal 
Force-, 



to design a thrust bearing which will operate satisfactorily at 
centrifugal-pump speeds. The multi-collar (Fig. 141), roller, 
and self -aligning ball (Fig. 142) types of bearings are used. In 
any event, the bearings of vertical pumps require considerable 
attention. 

137. When The Line Shaft Of 
A Vertical Pump Is Long, It Is 
Difficult To Keep The Motor, 
Line Shaft And Pump In Align- 
ment. When the line-shaft 
length exceeds about 30 to 40 
ft. a certain flexibility and the 
inevitable misalignment in an 
installation of this sort must be 
provided for. This necessitates 
the installation of several thrust 
bearings between the motor and 
the pump with a flexible coup- 
ling immediately above each 
thrust bearing. A guide bear- 
ing should be placed on each side 
of and close to each flexible coup- 
ling. The maximum distance 
between guide bearings should not exceed about 6 ft. 

138. The Performance Characteristics Of A Centrifugal 
Pump For Various Conditions Of Operation should be known 
before it is placed in any given installation. The factors which 
determine the performance characteristics of a centrifugal 
pump are principally : (1) The quantity of water delivered. (2) 
The efficiency. (3) The horse power input at each of several 
different heads. These data are usually supplied by the manu- 
facturer, but if they are not, they may be secured by test. . The 
two principal reasons for testing a pump are : (1) To determine 
its characteristics (Sec. 139). (2) To determine whether or 
not the manufacturer's guarantees have been fulfilled. 




Copper Bowl Ro- 
tates With ShaFt 



Fig. 143. — Guide Bearing For Verti 
cal Shaft Centrifugal Pumps. (Worth 
ington Pump And Machinery Corp.) 



Note. — A Centrifugal Pump May Be Tested As Follows: The 
pump which is to be tested is directly connected to a direct-current, 
variable-speed electric motor, M (Fig. 144) of known efficiency. A volt- 
meter, V ; and & n ammeter, A, are connected in tjtie motor circuit, A 



Sec. 138] CENTRIFUGAL AND ROTARY PUMPS 



123 



pressure gage, P, is connected into the discharge pipe. A vacuum gage, 
S, is connected into the suction pipe. The quantity of water discharged 
may be measured either by means of a calibrated nozzle placed on the 
end of the discharge pipe, or by a water meter, W. Or, if the pump is of 
small capacity, the water may be discharged into a suitable container, R, 
and measured directly. The gage readings of S and P should be com- 
bined and converted to head in feet (Sees. 5 and 38), which will be the 
total head pumped against if the discharge- and suction-pipes are of the 
same diameter. 



/Suction Gage 



AS Pressure. 
<® Gac,e-£ 

Pump->C\ 



A/^w— 



Source of D. C. Supply 



Voltmeter 

-Discharge Nozzle 

v -- x Discharge Reservoir 

\ RT 




„-;==*,•• Sill.' /,'/. -> ,. -'Ai'V-'i;- "*« w » J.*'*£ -- v •».■£ 

■•^ii.icr:-' 

Fig. 144. — An Arrangement Which May Be Used In Testing A Centrifugal Pump. 



The pump is primed and started. The speed must be maintained 
constant throughout the test. Simultaneous readings of S, P, A, V, and 
W are taken. S and P are converted into total head in feet (Sees. 5 and 
38). Then these formulas may be applied: 



(59) 
and 
(60) 



'bhp 



E P 



I X V X E, 
746 

L hT X V gm 
39.6 X Vbhp 



(horse power) 
(per cent.) 



Wherein: "P b h P = input to pump in horse power (also called brake horse 
power of pump). J = motor-current, in amperes, as read from the 
ammeter. V = motor-e.m.f., in volts, as read from voltmeter. V gm = 
quantity of water pumped, in gallons per minute, as determined from 
water meter. LhT = total head pumped against, in feet, as obtained 
from S and P. E m = efficiency of motor; at the given load, expressed 
as a decimal, as obtained from the motor efficiency graph. E p = effi- 
ciency of the pump, expressed in per cent. 



124 STEAM POWER PLANT AUXILIARIES [Drv. 4 

By applying the formulas for a certain discharge in gallons per minute, 
the head, the brake horse power, and the efficiency of the pump, when 
running at the given speed, are determined. The discharge is now varied 
by either opening or closing the gate-valve, G, and another set of readings 
is taken and the corresponding computations are made as described 
above. By opening or closing the gate-valve, the conditions should be 
varied from no discharge when G is closed, to practically no head when G 
is wide open. Several sets of readings should be taken, at fairly regular 
intervals of discharge in gallons per minute, over that discharge range 
which will be provided from gate-valve entirely closed to wide open. The 
test data should be plotted into a characteristic chart as will be described. 

Example. — A centrifugal pump (Fig. 144), which is undergoing test, 
is driven by a direct-connected, direct-current motor, at a constant speed 
of 1,700 r.p.m. A certain set of readings are as follows: V gm = 400 gal. 
per min.; S = 8.9 in. of mercury; P = 26 lb. per sq. in.; A = 36 amp.; 
V = 218 volts. Wha{ is the horse-power input to the pump, the head 
produced, and the efficiency of the pump in per cent., at this rate of 
discharge? 

Solution. — By note subjoined to Sec. 38, the suction pressure developed 
= (8.9 X 0.4914) = 4.37 lb. per sq. in. Since the water level is below 
the pump-center, S and P are combined by addition, or (4.37 + 26) = 
30.37 lb. per sq. in. By For. (1), the total head produced, Lht = (2.31 X 
30.37) = 70 ft. From the motor-characteristic chart, it is found that 
the efficiency of the motor at this load E m = 89 per cent. By For. (59), 
the horse-power input to the pump, ~Pbh P = I X FXE„-r 746 = 36 X 
218 X 0.89 -^ 746 = 9.37 h.p. By For. (60), the efficiency of the pump, 
E p = (L hT X V gm ) 4- (39.6 X V bhp ) = (70 X 400) -f- (39.6 X 9.37) =66.2 
per cent. 

Note. — A Centrifugal Pump Should Be Tested Under The Con- 
ditions To Which It Will Be Subjected When Installed. Thus 
a boiler-feed pump should be tested with water at the temperature of 
that which it will ultimately handle. 



139. A Chart Of The Characteristic Graphs Of A Centrifugal 
Pump may be plotted thus : First compute from the test data 
the head in feet, the brake horse power, and the efficiency, for 
each of the different rates of discharge. Then (Fig. 145) lay 
off, on the horizontal axis (on a sheet of cross-section paper), of 
the graph, to a convenient scale, the range of discharge values 
in gallons per minute. Next lay off, on the vertical axis, the 
range of values corresponding to the head in feet, the brake 
horse power, and the efficiency. Now plot the values: Lay 
off to scale, on the horizontal axis, distances equivalent in 
value to the different discharges in gallons per minute as 



Sec. 140] CENTRIFUGAL AND ROTARY PUMPS 



125 



taken from the test data. For each point thus obtained, 
locate new points in the body of the chart by laying off verti- 
cally, to scale, distances which are equivalent to the heads in 
feet for the discharge at each head. A smooth curve drawn 
through the points obtained as described, results in the head 
graph (Fig. 145). The brake horse power and the efficiency 
graphs are plotted in a similar manner. These three graphs 
are known as the characteristics of the pump. 




400 "W 1200 

Discharge In Gallons 



1600 . 2000 

Per Minute 



Fig. 145. — Typical Characteristics Of A Centrifugal Pump At Constant Speed. 

140. A Number Of Important Facts May Be Determined 
From The Characteristic Graphs (Fig. 145) of a pump which 
is operated at a given speed which are not apparent from the 
test data, such as: (1) The rate of discharge in gallons per 
minute when pumping against any head. (2) The efficiency 
of the pump at any discharge rate. (3) The horse power required 
to drive the pump when pumping water against any head. If 
(with a certain pump speed in r.p.m.) any one of the four items: 
the head pumped against, the efficiency, the brake horse 
power, or the discharge in gallons per minute, is known, then 
the other three can be determined directly from the graphs 
without further calculation. 

Explanation. — The highest point on the brake-horse-power graph 
(Fig. 145) is about 61 h.p. This indicates that a 60-h.p. motor would 
be suitable to drive the pump at any load without danger of motor- 
overload. It is also evident that the maximum efficiency is about 73 
per cent., and that when operating at this maximum emciencj^, about 
1,800 gal. per min. will be delivered against a 90-ft. head. When oper- 
ating under these conditions, the power required to drive the pump is 



126 



STEAM POWER PLANT AUXILIARIES 



[Div. 4 



about 59 h.p. Electric motors are usually designed to operate at their 
maximum efficiency at the rated full load. A 60-h.p. motor would, 
therefore, when driving the pump against a 90-ft. head, be operating at 
about its maximum efficiency. The maximum overall efficiency would 
be obtained with the pump direct connected to a 60-h.p. motor, when 
delivering 1,800 gal. per minute against a 90-ft. head. 

Note. — A pump, having a head graph similar to that of Fig. 145, 
has what is known as a rising characteristic. That is, beginning at 
shut-off, the head developed increases up to a certain point (about 
600 gal. per min. in this pump) and then decreases. A slightly rising 
characteristic is usually desirable. Note that after this 600-gal.-per-min. 
point is passed, that the horse-power input is increased, and that its 
efficiency increases up to a certain point, and then decreases. Study 
this graph of Fig. 145 to obtain a further understanding of the relations 
between head, efficiency, horse power, and discharge, in a centrifugal 
pump which is operating at a constant speed. 

141. Graphs Showing Typical Relations Between Head, 
Volume, R.P.M., And Efficiency, In Good Commercial 
Centrifugal Pumps are reproduced in Figs. 146, 147, 148 




60 d0 100 120 

Per Cent Of Rated Volume 

Fig. 146. — Relation Between Head, Volume, R.P.M., And Efficiency, In Good Com- 
mercial Centrifugal Pumps Which Operate Above 1,800 R.P.M. Against Heads Greater 
Than 50 Ft. 



and 149. (Marks' Mechanical Engineers' Handbook). 
There is no sharp division line between high head and low 
head. Low head is in these graphs, assumed to mean less 
than 50 ft. High head is assumed to mean above 50 ft. 
Low speed is up to 600 r.p.m.; moderate speed, from 600 to 
1,800 r.p.m.; high speed above 1,800 r.p.m. These graphs do 
not show the performance of any individual pump, but are 



Sec. 141] CENTRIFUGAL AND ROTARY PUMPS 



127 



the averages of data obtained from a large number of good 
commercial pumps, and show what may be reasonably ex- 
pected of the average pump. These curves are particularly 



JflO 


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10 



40 60 80 100 IZ0 

Per Cent. Of Rated Volume 



140 



160 



180 



Fig. 147. — Relation Between Head, Volume, R.P.M., And Efficiency, In Good Com- 
mercial Centrifugal Pumps Which Operate Between 600 And 1,800 R.P.M. Against 
Heads Greater Than 50 Ft. 

applicable to large-capacity pumps, as in the smaller pumps 
efficiency is likely to be sacrificed to decrease the first cost. 



..-Per Czntagz Of Rc/fzc/ R.RM. 




60 60 100 120 

Per Cent, Of Rated Volume 

Fig. 148. — Relation Between Head, Volume, R.P.M., And Efficiency, In Good 
Centrifugal Pumps Which Operate Betwen 600 And 1,800 R.P.M. Against Heads Less 
Than 50 Ft. 



Example. — A centrifugal pump has a normal rating of 8,500 gal. per 
min. when operating against an 80-ft. head at 1,700 r.p.m. About what 
efficiency should be expected when the pump is operating at its normal 
rating? If the speed is increased 5 per cent., what discharge rate should 
be expected if the head pumped against remains constant, and what 



128 



STEAM POWER PLANT AUXILIARIES 



[Div. 4 



should be the expected resulting efficiency? Solution. — Since a speed 
of 1,700 r.p.m. and a head of 80 ft. would classify this pump as moderate 
speed and high head, refer to the graphs of Fig. 147. The 100-per-cent. 
r.p.m. graph, the 100-per-cent. head graph, and the 100-per-cent. volume 
graph intersect at point A. It is found that an efficiency of about 80 
per cent, should be expected when the pump is operating at its normal 
rated load. The 105-per-cent. r.p.m. graph intersects the 100-per-cent. 
rated load graph at point B, which shows that about 112 per cent, dis- 
charge rate and about 74 per cent, efficiency may be expected with a 
5-per-cent. speed increase. 



180 

160 

1 '40 

n: 

."£ ioo 

S 80 

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v\ 


$ 


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v \ \ 


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v 


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w 








W 


A 


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zo 



4-0 60 80 I00 \Z0 I40 

Per cant. Of Rated Volume 



[60 



180 



Fig. 149. — Relation Between Head, Volume, R.P.M., And Efficiency In Good Cen- 
trifugal Pumps Which Operate Below 600 R.P.M. Against Heads Less Than 50 Ft. 



142. The Effect Of Changing The Conditions Under 
Which An Actual Centrifugal Pump Operates will now be 
considered. While the effect of changed operating conditions 
will always be determined primarily by the theoretical prin- 
ciples discussed in Sees. 118 to 121, these theoretical laws 
cannot, usually without modification, for reasons already 
suggested, be applied to an actual pump. In practice, the 
most feasible method of ascertaining the consequence of 
altered conditions is a graphic one. For this graphic method, 
a chart (Fig. 145), on which are shown the characteristic 
graphs for the pump under consideration, must be employed. 

Note. — How To Ascertain, From The Characteristic Graph, The 
Effect Of Changing Either The Head Or The Discharge, And 
The Corresponding Change In Efficiency and Horse-Power In- 
put, At A Constant Speed, has already been explained in Sec. 140. 
A method of obtaining the characteristic graphs at any desired speed (within 
reasonable limits) from the graphs for a given speed will be presented in 



Sec. 143] CENTRIFUGAL AND ROTARY PUMPS 



129 



the following Sec. Then, after having determined the characteristic 
graphs at any desired speed, the head, rate of discharge, efficiency, and 
brake horse power, can be ascertained at this desired speed. The 
charts for any pump at the rated speed may be obtained by test, or, 
usually, from its manufacturer by giving him a detailed description (all 
name plate data and serial number) of the pump under consideration. 

143. A Change In The Impeller Speed Of A Centrifugal 
Pump will (Sees. 118 to 121) affect the quantity of water 
delivered, the head produced, and the horse-power input. 
The formulas by which these variations are computed, for a 




Discharge In Gallons Pzr Minute 



Fig. 150. — Illustrating Method Of Determining Centrifugal Pump Characteristics At 
Any Desired Speed. 

theoretical installation without water-friction, are given in 
Sees. 118, 119, 120, and 121. This theoretical condition of a 
frictionless installation is very closely approximated in 
practice where a pump is delivering water to a stand-pipe 
through a short length of pipe without bends. However, in 
those installations wherein the friction head (Sec. 9) is large, 
these theoretical formulas cannot, without modification, be 
employed. Having the characteristic graph of a pump for a 
given speed, the method of obtaining the graphs for any other 
speed may be understood from a consideration of the following 
example : 



130 STEAM POWER PLANT AUXILIARIES [Dw. 4 

Example. — A 1,700-r.p.m. pump, which has a head graph as shown in 
Fig. 150 delivers 1,900 gal. per min. in a certain installation wherein the 
static head is 52 ft. and the friction head due to the piping is 50 ft. 
What quantity of water will be delivered and what head will be produced, 
provided the same piping is used, if the speed is decreased to 1,530 r.p.m.? 

Solution. — The friction head varies approximately as the square of 
the volume of water delivered. Therefore, at 950 gal. per min., the friction 
head = (950/1900) 2 X 50 = 12.5 ft. Take other values of discharge in 
gallons per minute and compute the corresponding friction heads in a 
similar manner. These values of friction head laid off vertically upward 
from the "52 ft. static" line (Fig. 150) on the corresponding discharge- 
rate lines result in the friction head graph. Next select points such as 
A, B, C, D, and E, on the 1700-r.p.m. head graph (Fig. 150) and deter- 
mine the head and discharge rate corresponding to each point selected. 
In Fig. 150, point C corresponds to 1,750 gal. per min. pumped against 
a 107-ft. head when the pump is running at 1,700 r.p.m. By For. (52), 
the quantity of water delivered at 1,530 r.p.m. = V gm 2 = (N2 X V gm i) ■*■ 
iVi = (1,530 X 1,750) -=- 1,700 = 1,575 gal. per min. By For. (53), the 
head produced at 1,530 r.p.m. = L h T% = (N 2 ■*■ JV1) 2 X L h n = (1,530 -s- 
1,700) 2 X 107 = 86.6 ft. Thus point C", which is a point on the 1,530- 
r.p.m. head graph, has a value of 1,575 gal. per min. at 86.6 ft. Simi- 
larly, determine the values of points A', B' , D', E', etc., which correspond 
to the values, of A, B, D, E, etc., and plot points A', B' , C, D', E' , etc., 
on the chart. A smooth curve drawn through A', B', C, D', etc., results 
in the 1,530-r.p.m. head graph (Fig. 150). The intersection of the 1,530 
r.p.m. head graph with the friction-head graph determines the quantity of 
water discharged and the head produced, which, in this case, is about 1,580- 
gal. per min. against about a 87-ft. head. 

Example -At what speed must the pump in the preceding example 
be driven to deliver 1,500 gal. per min.? 

Solution. — The l,500-gal.-per-min. line intersects the friction-head 
graph at point Y, and the 1,530-r.p.ra. head graph at X. Therefore, the 
speed required to deliver 1,500 gal. per min. = 1,530 — [{Distance XY -f- 
Distance XB) X (1,700 - 1,530)] = 1,530 - (0.24 X 170) = 1,530 - 
40 = 1,490 r.p.m., approximately. 

Note. — The Power Required To Drive A Centrifugal Pump 
At Any Speed, other than that upon which the available characteristic 
graphs are based, may be determined as follows: Suppose the chart is 
provided for the pump when running at 1,700 r.p.m. (Fig. 150) and that 
it is desired to determine the power required to drive the pump at 1,530 
r.p.m. From the available characteristic head graph construct the head 
graph for the desired speed, as explained above. When the pump is 
running at 1,530 r.p.m. and operating under the conditions which are 
represented by point B', it will have the same efficiency that it has when 
running at 1,700 r.p.m. under the conditions which are represented by 
point B; also the efficiency at C, D', E', etc., will be the same as that at 



Sec. 144] CENTRIFUGAL AND ROTARY PUMPS 131 

C, D, E, etc., respectively. Therefore, by projecting vertically down- 
ward from D, (Fig. 150) it is found that the pump, when operating at 
1,700 r.p.m. has an efficiency represented by S, of 72 per cent. Then 
locate point S' equivalent to 72 per cent, vertically downward from D'. 
S' is one of the points on the efficiency graph for 1,530 r.p.m. Other 
points are located in a similar manner and the 1,530-r.p.m. efficiency 
graph is drawn. From corresponding values of head, discharge rate, and 
efficiency, the brake horse power (power input to motor) can be computed 
by For. (60) and the graph can then be drawn, as explained in Sec. 139. 

144. The Methods Of Driving Centrifugal Pumps Are: 

(1) Belt or ropes. (2) Direct connected to an electric motor. 
(3) Direct connected to a steam or gasoline engine. (4) Direct 
connected or reduction-gear connected to a steam turbine. Each 
will be discussed: 

145. A Belt Drive For Centrifugal Pumps is better suited 
to those of small than to those of large capacity. It should be 
employed only when direct-connection is infeasible. When 
it is desired to use a belt drive, a pump which has a relatively 
low speed should be selected. In general, the belt speed 
should not be permitted to exceed about 4,500 ft. per min. 
The pulley centers should be located a reasonable distance 
apart, especially when there is much difference in the size 
of the driving pulley and the driven pulley. The tight side 
of the belt should, when possible, be underneath. 

Note. — When The Arc Of Belt-Contact Is Approximately 180 
Degrees, The Required Width Of A Single Belt To Drive A Cen- 
trifugal pump may be computed by the following formula: 

(61) L w = N xd (inches) 

Wherein: L w = width, in inches, of a igngle belt. Pbh P = maximum 
brake horse power required to drive thgpump. N = revolutions per 
minute at which the pump is to operate, fl = diameter, in inches, of the 
pulley on the pump shaft. To obtain the required width of a double 
belt, multiply the result obtained from For. (61) by 0.625. The pulley 
used on the pump shaft should have a face at least 2 in. wider than the 
belt. 

146. The Direct-Connected Motor -Driven Centrifugal 
Pump (Fig. 157) is one of the most satisfactory forms of 
centrifugal-pump installations. The principal reasons for 
this are: (1) Saving of floor space. (2) Reduction of power- 



132 STEAM POWER PLANT AUXILIARIES [Div. 4 

transmission losses. Since the centrifugal pump is a relatively 
high-speed machine, and as high-speed motors are cheaper 
than low-speed motors, a saving in the first-cost is obtained. 
By the use of a variable speed motor, various pumping condi- 
tions (Sees. 118 to 120) may be satisfied by the same unit. 

Note. — The Electric Motor As A Drive For Centrifugal Pumps 
Has Decided Advantages in isolated installations or where no facilities 
are at hand for utilizing the heat available in the exhaust steam. 

Note. — Direct-Current Motors find application for installations 
where only direct current is available or where adjustment of speed is 
necessary. The direct-current motor has the further advantage in that 
it can be designed for any definite speed. Where the voltage is constant 
either shunt-wound or the compound-wound, direct-current motor can 
be used with success. Where the voltage is variable, as in some tem- 
porary installation, particularly when fed from an electric railway cir- 
cuit (see Sec. 173), a compound direct-current motor should be used. 
It is generally recommended that, when direct-current motors are used, 
the discharge gate valve be closed in starting. This procedure should 
especially be followed with shunt-wound motors. Proper ventilation 
must not be overlooked in motor-driven installations. 

Note. — Motor-Driven Centrifugal Pumps Are Usually De- 
signed To Operate At Speeds Of About 1,100, 1,200, 1,700, And 1,800 
R.p.m., since these are the more usual "synchronous" speeds of alter- 
nating-current motors. The synchronous speed of any QO-cycle alternating- 
current motor = 7,200 -5- number of poles. The actual full-load induction- 
motor speed will be about 5 per cent, less than the synchronous speed. 
Direct-current motors are often designed to run at these speeds. This 
renders a pump which is designed to operate at one of the above speeds 
suitable for either direct- or alternating-current-motor drive. 

Note. — The Power-Factor-Correcting Ability Of The Syn- 
chronous Motor Is Increasing The Demand For Direct-Connec- 
ted, Synchronous-Motor-Driven Centrifugal Pumps. If, however, 
the brake horse power at shut-off is greater than about 35 per cent, of 
the full-load brake horse powder, difficulty is likely to be experienced in 
getting the motor to pull into synchronism. 

Note. — The Squirrel-Cage Alternating-Current Induction 
Motor Is Well Adapted To Centrifugal-Pump Drives (R. A. Fiske) 
because of the simplicity of the motor and its control. The first cost is 
generally less than that of a motor of the slip-ring type. Due to the 
squirrel-cage motor's characteristic of low starting torque, a valve 
should, where one of these motors is used, be placed in the discharge line 
to minimize the load on the pump during the starting period. 

Note. — Slip-Ring Induction Motors are preferable for centrifugal 
pumps of the larger capacities because of their ability to start smoothly 
against great torques without taking excessive currents from the line. 



Sec. 147] CENTRIFUGAL AND ROTARY PUMPS 133 

147. A Steam Or Gasoline Engine, Direct Connected To A 
Centrifugal Pump constitutes an economical method of opera- 
tion. The speed of ordinary reciprocating machinery is, 
however , relatively low. Consequently this method of 
drive is only suitable for the low- and medium-head pumps. 

148. The Direct- Or Gear-Connected Steam-Turbine 
Centrifugal -Pump Drive is rapidly gaining in popularity. 
Since both the steam turbine and the centrifugal pump are 
inherently high speed machines, they are admirably suited to 
each other. The steam-turbine-driven centrifugal pump is 
even more flexible as to speed variation than is a motor-driven 
pump. By the installation of suitable governors, which are 
actuated by the pump discharge-pressure, control is obtained 
whereby the turbine speed is automatically adjusted so that 
the head produced by the pump remains constant over a 
range of from J4 to full pump-capacity. The maximum eco- 
nomical speed for large-capacity pumps operating against low 
heads is usually lower than the minimum economical turbine 
speed. Hence in such installations the pump is connected 
to the turbine through specially-designed reduction gears. 
This enables both turbine and pump to be driven at their most 
economical speed. There is but little power-transmission 
loss (about 2 per cent.) through a reduction gear of the double 
helical type. 

Note. — The Steam Turbine When Exhausting Into A Vacuum 
Affords A Very Economical Drive. (R. A. Fiske.) In such units 
the turbine may exhaust into a condenser (Div. 9) serving one or several 
of the main generating units. Or the exhaust may be used to advantage 
in feed- water heaters (see Div. 7). Where economizers are installed and 
where there would otherwise be an excess of auxiliary exhaust, low-pres- 
sure turbines could be used as drivers, the low-pressure steam being 
derived from other auxiliaries or from the intermediate receivers of 
the main engines or turbines. The turbine can also be arranged to ex- 
haust into the intermediate stages of the main generating units. For 
the larger units, it may prove advantageous to use a low-level jet con- 
denser (Sec. 336) taking the condensing water from the discharge side 
of the pump and returning the waste water to the suction well. 

149. A Flexible Coupling Should Always Be Used To Direct- 
Connect A Centrifugal Pump To Its Motive Power. — Usually, 
the pump and the driving unit have two main bearings each. 



134 



STEAM POWER PLANT AUXILIARIES 



[Div. 4 



If a rigid flange-coupling is used, the driving-unit shaft and 
the pump shaft become, in effect, a solid continuous shaft, and 
it is practically impossible to align four bearings so that each 
will function properly at centrifugal-pump speeds. A flexible 
coupling (Fig. 151) will compensate for slight inaccuracies in 
alignment, and also reduce vibration. In a gear-connected 

steam-turbine drive, a suit- 
able flexible coupling should 
be used on each side of the 
gear. 

Note. — It is generally con- 
ceded that flexible couplings 
may not prove entirely "flex- 
ible" on high-speed shafts. 
Therefore, a rigid baseplate 
extending under pump, driving 
unit and gears, should always 
be provided to maintain the 
shafts of the two units in 
good alignment, especially 
where the pump is driven at 
speeds above 1,500 r.p.m. 




Fig. 151. — Flexible Coupling For Direct- 
Connecting A Centrifugal Pump To Its 
Driving Unit. 



150. The Advantages Of The Centrifugal Pump (R. A.Fiske, 
The Centrifugal Pump, Power Plant Engineering, February 
15, 1921) are: (1) But one moving part. (2) No valves or pistons 
to be kept in order. (3) Uniform pressure and flow of water. (4) 
Freedom from shock. (5) Compactness. (6) Simplicity of 
design. (7) Simple to operate and repair. (8) High rotative 
speed, allowing direct connection to electric motors and steam 
turbines. (9) In case of stoppage of delivery, the pressure cannot 
build up beyond predetermined working limits. (10) Low first 
cost. (11) Low rate of depreciation. 

151. The Disadvantages Of The Centrifugal Pump are: (1) 
The rate of flow cannot be efficiently regulated for wide ranges 
in duty. (2) The efficiency is not usually as high as the best 
grade of piston pump. (3) Direct connection to low-speed engines 
cannot be made when operating on high lifts. 

152. A Comparison Between Centrifugal Pumps And Recip- 
rocating Pumps will explain the increasing demand for the 
centrifugal pump. The centrifugal pump is, in general, su- 



Sec. 153] CENTRIFUGAL AND ROTARY PUMPS 135 

perior to the reciprocating pump in simplicity, reliability, 
ease of operation, durability, space occupied, and frequently 
in over-all efficiency. It has a more uniform discharge pressure 
than has the displacement pump, it vibrates less and does not 
require as heavy a foundation. Except for very small ca- 
pacities, the average first cost of centrifugal pumps is about 
}i of that of reciprocating pumps. The centrifugal pump is 
capable of handling water which contains gravel, sand, and 
if suitably designed even fair-sized rocks. This is impossible 
with the other type. The inherent characteristics of the 
centrifugal pump render it unsuitable for services which re- 
quire a very positive control of capacity and head. The 
centrifugal pump is not well adapted to services which require 
a high suction-lift (Sec. 88), nor for pumping small quantities 
of water against high heads. 

Note. — The Centrifugal Pump When Designed For A Certain 
Capacity And Head Cannot Be Used, Without Great Loss In 
Efficiency, At Any Other Capacity Or Head. It is not as flexible 
in this respect as is the reciprocating pump, which can be used under 
widely different conditions without any great sacrifice in efficiency. 

153. The More Common Services To Which Centrifugal 
Pumps Are Applicable are: (1) Sewage pumping plants. (2) 
Dry dock pumps. (3) Irrigation and drainage. (4) Condenser 
circulating pumps. (5) Municipal water works. (6) Hydraulic 
elevators. (7) Mine drainage and hydraulic mining. (8) 
Fire pumps. (9) Boiler-feed service. Pumps of the volute 
type are more generally used for the four services which are 
first mentioned, and those of the turbine type for the five 
services last mentioned. 

Note. — Certain of these services will be discussed in following Sections 
but the scope of this book does not permit a detailed discussion of all. 

Note. — About the only services for which centrifugal pumps cannot 
be applied are high-pressure-hydraulic-press and deep-well service. 

154. The Centrifugal Pump Is In Almost Universal Use 
For Circulating The Condensing Water (Div. 9) in modern 
power-plant installations. It is applicable to jet-, barometric-, 
and surface-condenser service. The steam-turbine drive is 
particularly applicable to barometric-condenser service (Sec. 



136 STEAM POWER PLANT A UXILI ARIES [Div. 4 

339) wherein a higher head is required at starting than under 
normal operating conditions. The turbine can be speeded 
up to produce the desired initial head. Then when the 
vacuum becomes established, the speed can be reduced to 
that required for normal operating conditions. Where the 
static head varies, as it is likely to do where the water is 
pumped from a river, variable speed operation is especially 
desirable. 

155. Boiler Feeding By Centrifugal Pumps (Div. 6) com- 
prises one of the most desirable applications for this type of 
pump. It is, however, not to be recommended for small-plant 
service. The unit occupies but little space, and requires only 
a light foundation. It can be started when cold and put into 
service in a very short time. The absence of vibration is an 
important feature, Excessive vibration in a boiler-feeding 
apparatus will open the pipe joints. 

Note. — A series of tests which were made by the Terry Steam Turbine 
Co. show an average of 62.4 lb. of steam per horse-power hour for the 
steam-turbine-driven centrifugal pump as compared with 91.9 lb. for 
the reciprocating pump. The tests were made on boiler-feed service at 
a discharge rate of 300 gal. per min. against a 200-lb. total net head. 

156. The Selection Of A Centrifugal Pump For Boiler- 
Feed Service (See Div. 6) requires, primarily, a consideration 
of: (1) Capacity. (2) Discharge or boiler pressure. (3) 
Location with respect to the feed water. (4) Load factor of the 
plant. A boiler-feed pump must always be designed for excess 
capacity over that required for the rated boiler horse power 
which is to be served (Sec. 225). Where high peak loads are 
carried for short periods, the installation of duplicate units is 
advisable. One can then be operated under normal loads, 
and both during the peak-load period. When no economizers 
are used the discharge pressure for a centrifugal boiler-feed 
pump should exceed the boiler pressure by about 25 lb. per 
sq. in. When economizers are used, the excess should be 
from 35 to 50 lb. per sq. in. 

157. Where A Centrifugal Pump Is Used For Boiler Feed- 
ing In Connection With A Feed -Water Heater, the hot water 
should flow to the pump under a positive head. Any suction 
pull which is exerted on water will cause a reduction in the 



Sec. 158] CENTRIFUGAL AND ROTARY PUMPS 



137 



absolute pressure and a consequent lowering of the boiling 
point and the water will tend to vaporize. If the total pressure 
exerted at the pump-suction nozzle by the weight of the hot- 
water column is insufficient to overcome this tendency, vapor 
will be formed in the pump, and the pump will become vapor 
bound. This will reduce the pump-capacity, or it may cause 
the pump to entirely stop delivering water. If the tempera- 
ture of the water to be pumped exceeds 120 deg. fahr. the 
installation should be so arranged that the suction head (Fig. 
153) is positive. For temperatures above 120 deg. fahr., there 









































12 




















































































































































s 

5 










































































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s 






































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130 



140 , 150 160 170 180 190 200 

Approximate T&mperotture In Decjrws Fahrenheit 



£10 



220 



Fig. 152. — Graph Showing Effective Head Required At Pump-Suction Inlet To Suc- 
cessfully Handle Hot Water With Centrifugal Pumps. (Based On Data In Alberger 
Pump And Condenser Company's Catalog.) 



should be an effective head on the center of the pump equiva- 
lent to that shown by the graph in Fig. 152. 

Note. — Centrifugal Pumps Will Deliver Water At Somewhat 
Higher Temperatures under the corresponding heads on the suction 
nozzle than those given by the graph of Fig. 152, but the efficiency will 
thereby be materially decreased. 

158. The Data Which Should Be Furnished The Pump 
Manufacturer When Requesting Quotations (R. A. Fiske) 
are: (1) Capacity of pump — gallons per minute. (2) Total 



138 STEAM POWER PLANT AUXILIARIES [Div. 4 

lift, including discharge and suction head as well as pipe 
friction. (3) Variations in lift, both suction and discharge. 
(4) Type — horizontal or vertical. (5) Quality of liquid, 
fresh water, gritty or solids in suspension. (6) Temperature 
of liquid. (7) Specific gravity of liquid. (8) Service, water 
works, irrigation, boiler-feed or what. (9) Motive power 
to be used. 

Note. — In The Selection Of A Pumping Unit It Is Always Best 
To Obtain Bids From Several Manufacturers; study over the bids 
carefully; tabulate them in so far as the primary features are concerned; 
make a careful comparison of details as to ease of dismantling, method 
of lubrication, size and construction of bearings, materials used, and the 
general ruggedness and serviceability of the pump as a unit. Then pur- 
chase the unit which, on an annual cost basis (see Sec. 366 on Condensers) 
will probably be the most economical. 

Note. — Properties Sometimes Disregarded When Selecting A 
Centrifugal Pump are: (1) Two single-stage pumps each capable of 
delivering 500 gal. per min. at 75 ft. head, connected in series will deliver 
500 gal. per min. at 150 ft. head. The same pumps connected in parallel 
will deliver 1,000 gal. per min. at 75 ft. head. (2) A centrifugal pump 
never loses any head that may be received by it at the suction chamber. 
For example : if a pump capable of delivering water ai^ 200 ft. head receives 
it at 100 ft. head it will discharge the water at 300 ft. head. (3) When 
two or more equally rated centrifugal pumps discharge into a common 
main their characteristics should be the same; otherwise, under certain 
conditions of head, they may not give equal delivery and there is the 
possibility of one pump cutting out altogether. 

Note. — It Often Happens That The Driving Unit Selected Is 
Not Manufactured By The Pump Builder. The builder will always 
quote on any standard driving unit which may be required. This raises 
the question as to whether or not the purchaser can effect a saving by 
buying the driving unit direct from the manufacturers and having it 
shipped to the pump manufacturer and assembled to the pump by the 
latter, thus eliminating the carrying charges expected by the pump 
builder. The answer to this is, avoid divided responsibility: Have the 
pump builder furnish the complete unit, and then hold him alone re- 
sponsible for its effectiveness. 

159. Proper Installation Of A Centrifugal Pump requires, 
first of all, that the foundation be built sufficiently massive 
and rigid to avoid excessive vibration. Vibration of the 
stationary details of a centrifugal pump tends to produce 
losses, due to excessive mechanical friction and to leakage 



Sec. 160] CENTRIFUGAL AND ROTARY PUMPS 



139 



through loosened joints. The pump should be set as close as 
possible to the level of the water at the source of supply. It 
will, in every case, be advantageous if the pump can be set 
(Fig. 153) below the level of the suction supply, so that the 
water may flow to it by gravity. Where water at a higher 
temperature than about 120 deg. fahr. is to be handled, this 
provision is practically imperative. The pump should be so 
located as to avoid elbows, bends or other sources of friction 
in the suction line. Where a suction-lift cannot be avoided 
altogether, it should, if possible, be less than 15 ft. 



Note. — A Suction-Lift Of More 
Than 15 Ft., including the head due 
to friction in the pipe, should not be 
attempted with pumps of larger size 
than 8-in. discharge. With pumps of 
smaller size, however, it may be pos- 



Discharg/e 
,-Pump Casing/ 




Suction , 
Reservoir-' 



Suction Nozxlz 




Fig. 153. — Centrifugal Pump Set Below 
Level Of Suction Supply. 



Fig. 154. 



-A Right-Hand Centrifugal 
Pump. 



sible to operate with satisfactory results (should the conditions of 
installation require) with a total suction-head up to 20 ft. 

160. The Direction Of Rotation Of A Centrifugal Pump 

should be determined with a view to convenience and adapta- 
bility in the installation of the pump. The main factor to be 
considered is that the rotation shall be such that the suction 
and discharge nozzles will be so disposed as to permit the lines 
of suction and discharge piping to be run as nearly straight 
as possible. 

Note. — A Right-Hand Centrifugal Pump (Fig. 154) is one which, 
when viewed by an observer from the motive-power end, rotates clock- 



140 



STEAM POWER PLANT AUXILIARIES 



[Div. 4 



wise. A left-hand centrifugal pump (Fig. 155) is one which, from the 
same viewpoint, rotates counter-clockwise. 

Note. — Centrifugal Pumps Are Constructed With The Volutes 
Or Casings So Arranged (Fig. 156) As To Secure Considerable 



^Oi 



Dirzction Of Rotation*. 




M 








<D 




I-L&ft Homo! 













W//////V//////////////, 



Fig. 155.— A Left-Hand Centrifu- 
gal Pump. 



2- Right Homol 

Fig. 156. — Diagram Showing Various Dis- 
charge Positions Of Centrifugal-Pump Dis- 
charge-Nozzles. 



Suction 
Nozzle--. 



■■—Discharge Nozzle 
Driving Motor--'"'^^ 
Shaft Coupling^ 



Centrifugal) 
Pump— ' 



5 



Diversity in the directions of the discharge tangents. This renders it 
possible, in most cases, to select a construction which will permit of a 
straight run of discharge piping directly from the pump. 

161. The Foundation 
For A Centrifugal Pump 

should be made of con- 
crete or brick (see the 
author's Machinery 
Foundations). It should 
be built up to within about 
0.75 in. of the level (Fig. 
157) at which the pump is 
to stand. This is to pro- 
vide space for leveling and 
grouting the bed-plate of 
the pump. 

Note. — The Foundation Bolts, For Holding Down A Centrifu- 
gal Pump, should have at least 6 in. of thread at their upper ends. 



Lining- Up} 


5»,| 




II J, 


Wedge- - 


\\u 




I 


' ;t , GV^S ^'u'^dH 


1 V - 


Casing—. 


1 




'vJ 



■-Foundation Bolts- 

Fig. 157. — Section Of Foundation Of Centrifu- 
gal Pump. 



Sec. 162] CENTRIFUGAL AND ROTARY PUMPS 



141 



Thimbles (Fig. 157) made of standard iron pipe should be set in the ma- 
sonry so as to form casings for the foundation bolts. The thimbles should 
be large enough to give about 0.5 in. clearance around the bolts. 

162. To Level A Centrifugal Pump, after the pump has 
been set on the foundation with the bolts projecting through 
the holes in the bed-plate, two or more iron wedges (Fig. 157) 
should be inserted under each of the four edges of the plate. 
The pump may then be wedged up to the proper level. The 
packing in the stuffing-boxes should then be loosened and the 
revolving parts turned by hand to test their freedom of 
movement. 

Note. — The Bed-Plate Of A Centrifugal Pump Should Be 
Grouted (Fig. 157) with liquid cement poured into the space between 
the top of the foundation and the bed-plate. When the grouting has had 
time to thoroughly set and harden, the foundation-bolt nuts may be 
tightened down. Care should be exercised not to draw the nuts so tightly 
as to spring the bed-plate. This, however, is not likely to occur if the 
space beneath the plate is completely filled with the grout. 

163. The Suction Piping Of A Centrifugal Pump should in 
no case be of smaller size than the suction orifice in the pump 
casing. If the suction lift exceeds 15 ft., however, it is 



Discharge Nozzle^. ,-Increaser 



..-Garfe 
Valve 




5ucfhn P/pefOf Larger Size Than 
Suction Nozzle Of Purnp)^ 
\ 

3 f 

:-w- v ^» "j^V.- ^y ~ ^ ^ ~Z?~ > ~& r 
Foof Vaive^ 



100 Or More 



Sucfion Nozzle 



At Least 3 Ft.-.: 




Fig. 158. — Long-Draft Suction Piping For Centrifugal Pump. (The " increaser" 
should be arranged as in Fig. 162 instead of as here shown. Also, the suction pipe 
should pitch downward away from the pump.) 

generally advisable to use piping one size larger than the 
suction opening. And if the horizontal distance over which 
the suction supply must be conveyed is very great, say in 
excess of 100 ft. (Fig. 158), piping at least two sizes larger than 
the inlet in the casing should be used. This is to avoid 
excessive friction. Where the water must be lifted, the suc- 
tion pipe should extend (Fig. 158) at least 3 ft. below the 
level of the water in the suction-well, or other source of 



142 



STEAM POWER PLANT AUXILIARIES 



[Div. 4 



supply. This is to prevent air being drawn into the pump. 
Scrupulous care should be exercised, in laying out the piping, 
to avoid pockets for the accumulation of air. Where there is 
any suction lift, all portions of the piping should pitch down- 
ward toward the source of supply. In such cases no consider- 
able length of the piping should run horizontally, and no 
part of it, of whatever length, should be allowed to pitch 
downward toward the pump. If the water flows to the pump 
by gravity, or is supplied under the pressure of a pumping 
system as from street-mains, a gate valve should be inserted in 
the suction line close to the pump. This is for convenience in 




-Valve -Body '-Sfrainzr 
Fig. 159.— Foot Valve With Stainer. 




'Hmgeo/ Rubber- 
Facec/ Va/vzs. 



Fig. 160.— A Foot-Valve 
sembled. (Strainer shown 
from seat-flange.) 



Disas- 
broken 



case it might be necessary to disconnect the pump for repairs. 
If the pump is required to lift its suction supply, a foot-valve 
(Figs. 159 and 160) should be connected to the inlet orifice of 
the suction pipe. All right-angled turns in the piping should 
be made with long-radius bends. 

Note. — Where A Number Of Centrifugal Pumps Are Required 
To Take Their Suction From A Common Source each pump should 
(Fig. 161) have an independent suction line. 

Note. — A Foot Valve (Figs. 159 and 160) is merely a check valve, 
provided with a strainer, which is arranged for attachment to the lower 
end of a suction pipe. 

Note. — It Is Inadvisable To Terminate The Suction Line In An 
Elbow Connected Directly To The Suction Nozzle of the pump. 
An elbow so connected causes a whirling motion of the entering water. 



Sec. 164] CENTRIFUGAL AND ROTARY PUMPS 



143 



This tends to produce an irregular flow. Also, if the pump is of the 
double-suction type (Fig. 130) it tends to cause a condition of unbalance 
between the pressures on the two sides of the double impeller. 

Where it is necessary to insert an elbow at the pump end of the suction 
line, a short length of straight pipe should intervene between the elbow 



{DrivingBelt 



Suction Pipe 
QfNo.Z Pump, 



Suction 

Pipe Of 

t Mo.I Pump 




Fig. 161. — Centrifugal Pumps Drawing Through Independent Suction Lines From 

Common Source. 

and the suction nozzle of the pump. If the suction line is of larger size 
than the nozzle, then a tapered reducer may be used. The reducer 
should, however, be of eccentric form (Fig. 162) so as to avoid the air 
pocket which (Fig. 163) would result if a reducer with concentric ends 
were used. 



.-Discharge Nozzle 

.■Suction Nozzle 
• .-Eccentric Reducer 
. Elbow-,, 



■ ■•-■:■.:■. I suction Line---. 
■•J:j? C o 'n c r « e t <£££>•;•* 




Discharge Nozzle-, 
Suction Nozzle^ 
Air-Pocket-, j 

fElbow 





-4 — ■ -Suction Line 



UO 



;,j'Concret Q.-S^-? < i 



Fig. 162. — Showing How Enlarged 
Suction Line Should Be Connected To 
Centrifugal Pump. 



Fig. 163. — Showing How Concentric 
Reducer In Centrifugal-Pump Suction 
Line Forms An Air-Pocket. 



164. The Suction Pipe Of A Centrifugal Pump For Drawing 
Water From A Driven Well (Fig. 164) should, if the well is 
deep enough, run down inside the well-casing to a depth of 
about 25 ft. The annular space between the suction pipe and 
the casing should be sufficient to permit free access of air to the 
surface of the water in the well. This space should be left 
uncovered at the top of the casing. 

Note. — A Single Centrifugal Pump Should Not Be Required 
To Lift Water From A Battery Of Wells Or Sumps. A separate 



144 



STEAM POWER PLANT AUXILIARIES 



[Div. 4 



pump should be connected to each well. Where the suction piping of a 
single pump is connected to more than one well, the pump will tend to 
draw the larger part of its supply from the well or wells nearest to it. 



Guide 
Bearings 




I-Front Elevation Of 
Steel Supporting Frame 



I 

n-Side Elevation Of 



ll-Diae Elevation ut 
Steel Supporting Frame 

Fig. 164. — Centrifugal Pump Drawing Water From A Driven Well. 



One of the wells may thus be pumped down to the level of the inlets to 
the piping. When this occurs, air will enter the piping and break the 
suction from the other wells. 



Sec. 165] CENTRIFUGAL AND ROTARY PUMPS 



145 



Note. — A By-Pass Between The Suction- and The Discharge- 
Piping Of A Centrifugal Pump (Fig. 165) may often be useful as a 
means of preventing the pump from losing its suction, due to low water. 
This may occur where the pump is used for drawing water from a sump. 
By adjustment of the valves A and B the water in the sump can generally 
be kept at any desired level. 

165. An Air -Chamber In The Suction Line Of A Centrifugal 
Pump (Fig. 166) may be necessary where the water that is to 
be pumped is so impregnated with air that the suction lift 
cannot otherwise be steadily maintained. By running a pipe 
from the top of the air-chamber to the vacuum pump of a 



^ -Discharge 
Pipe 

Driving Be/r-^ 
/ 




'-■■Vacuum Pipe (Connected To A 

Vacuum System Or To An Air 

Pump) 

A/'r Chamber-.^^ 
Discharge Pipe-^^ 
.-Driving Motor 
\ Centrifugai Pump 




Fig. 165. — Centrifugal Pump With By- 
Pass. 



Fig. 166. — Centrifugal-Pump Suction 
Line Equipped With Air-Chamber. 



condenser or of a heating system, or to any system of piping 
in which a vacuum is maintained, the air can be removed as 
fast as it separates from the suction water. 

Note. — The Vacuum Pipe leading from the air-chamber in a cen- 
trifugal-pump suction line should extend to a vertical height of at least 
34 ft. above the level of the suction nozzle of the pump. This is to 
insure that no water will pass from the suction line into the vacuum 
system to which the air-chamber is piped. 

166. The Discharge Piping Of A Centrifugal Pump should 
be laid out with a special view to minimizing pipe friction. 
Unnecessary or avoidable frictional resistance in the piping 
means a wasteful expenditure of power in driving the pump. 
The piping should never be of smaller size than the discharge 
nozzle of the pump. But where pipe friction, due either to 
excessive length of the line or to unavoidable turns therein, is 

10 



146 



STEAM POWER PLANT AUXILIARIES 



[Div. 4 



a considerable item, it is usually advisable to minimize it by 
using pipe of a larger size than the pump connections. 

Note. — The Flow Velocities And Corresponding Frictional 
Resistances Of Systems Of Discharge Piping may be computed, and 
adequate pipe sizes selected by using the values given in Table 14 or 15. 

167. The Discharge Pipe Of A Centrifugal Pump Should 
Contain A Check -Valve (Fig. 167), located as closely as 
possible to the pump. The function of the check-valve is to 
protect the pump-casing from breakage, due to waterhammer 

that might occur in the dis- 
charge line. Waterhammer is 
particularly liable to damage a 
centrifugal pump which is un- 
protected by a check valve, if 
a float valve is attached to the 
suction line. 



Note. — A Gate Valve Should Be 
Installed In Addition To the check- 
valve in the discharge line of a cen- 
trifugal pump. The check-valve 
should be connected between the gate 
valve and the pump discharge nozzle. 
The function of the gate valve is to 
afford a means for controlling the dis- 
charge from the pump. It also serves 
to isolate the check-valve from the 
discharge piping in the event of repair 
or inspection of the check-valve be- 
coming necessary. 




Fig. 167. — Priming Ejector For 
Use With Pump Unprovided With 
Foot Valve. (Valve A is first opened. 
The steam valve B is then opened. 
When water begins to flow from the 
ejector nozzle, C, the pump is primed. 
Valves A and B may then be closed and 
the pump started.) 



168. Centrifugal Pumps Re- 
quire Priming. — That is, the 
casing of the pump must be 
filled with water before the im- 
peller is set in motion. Where the pump is below the water 
level of the source of supply (Fig. 153) it may be primed 
simply by opening the gate valve in the suction pipe. Where 
a foot-valve is used (Fig. 168) and the discharge pipe is con- 
nected to an overhead tank or reservoir, a by-pass (Fig. 168) 
may be connected between the discharge pipe and the suction 
pipe to compensate for leakage through the foot-valve while 



Sec. 169] CENTRIFUGAL AND ROTARY PUMPS 



147 



the pump is shut down, 
of water at all times. 



The casing may thus be kept full 



Note. — A Centrifugal Pump Should Not Be Run When Its 
Casing Is Empty. Certain of the 
interior parts are lubricated only by 
the water which passes through the 
pump. Running the pump dry is 
ruinous to these. 



Priming By-Pass— 

Discharge Pipe __ 

Driving Belf-^ 




169. There Are Several 
Methods Of Priming Centrifu- 
gal Pumps. — A vacuum pump 
may be connected (Fig. 169) to 
the opening in the top of the 
casing. Then with a check-valve 
in the discharge pipe or with the 
gate valve in the discharge pipe 
closed, the air may be exhausted 
from the casing so that the 
water will rise and fill the casing through the suction pipe. 
The same effect may be produced by running a pipe from the 

..-Check Valve 

rPitcher-Spout 

Pump 

.--Handle 



foundation 



Fig. 168. — Showing How The Dis- 
charge Pipe Of A Centrifugal Pump 
Should Be Valved. 




Check 

Valve ■' 



Fig. 169. — Centrifugal Pump Ar- 
ranged For Priming By Means Of A 
Vacuum Pump. 




Fig. 170. — Priming-Pump Mounted 
On Casing Of Centrifugal Pump. 
(Valve A is first opened. Handle, H, 
is then worked until water appears at 
Spout B. Valve, A, may then be closed 
and the pump started.) 



top opening in the casing to a steam-condensing system, or to 
any system of piping in which a vacuum exists. In some 



148 



STEAM POWER PLANT AUXILIARIES 



[Div. 4 



situations it may be convenient to fill the pump casing directly 
from the city water mains, or from an elevated tank in a 
house or factory water supply system. Or a hand-pump, 
mounted either on the pump casing (Fig. 170) or on the wall 
nearby (Fig. 171) may serve as a priming apparatus. The 
siphoning effect of a jet of steam, compressed air or water is 
also commonly utilized in the priming of centrifugal pumps. 



T I,L 


I i. ' i '- 


tt^~?T^ 


1 — i i i ' , i 


^ 




iC/sfernL-j^^x i ' i ' i ' i 


V- 1 


if-H-H- 


TS T 7^^m\h- L r L T J - 


3 3"r" 


"=V^ 


; ■ ; i , i ; ^jp^x^ 




-Nozzlz^ 


s^SS 




pS 




^D 


fc(?cy 
ririnof L 




y Handle ■£ 

rifugal-Xz 
Cas/ngzx~ 



Fig. 171. — Priming-Pump Mounted On 
Wall. (Valve, A, is first opened. 
Handle, H, is then worked until water 
appears at nozzle, N. Valve, A, may- 
then be closed and the pump started.) 



Steam Or . 
Compressed Air 
5upp/y-P/pe 

Ejecfor- 




Fig. 172. — Priming Ejector For Use 
With Pump Provided With Foot Valve. 
(The nozzle- valve, A, is first opened. The 
steam valve, B, is then opened. When 
water begins to flow from the bleeder, C, 
the pump is primed. Valves A, B, and C, 
may then be closed and the pump started.) 



The device which embodies this principle is called a priming 
ejector. Fig. 167. 

170. Where No Foot-Valve Is Used (Fig. 167) the priming 
ejector should be so arranged that the current of steam or 
compressed air will draw the air out of the pump casing. The 
water will then rise through the suction pipe of the pump. 
Where a foot-valve is used (Fig. 172) the ejector should be 



Sec. 171] CENTRIFUGAL AND ROTARY PUMPS 



149 



so arranged that the current of steam or compressed air will 
draw the water up through the suction pipe of the ejector 
and discharge it into the pump casing. 

Note. — With A Low Suction Lift, 6 ft. or less, a centrifugal pump 
can be primed, where the discharge pipe is filled with water and no foot- 
valve has been provided, by first starting it in motion and then letting 
water flow into the casing through the discharge pipe. But this is a 
very objectionable method and should not be attempted where other 
means are available. When a centrifugal pump is primed in this manner, 
the load is suddenly thrown on while the apparatus is rotating at a high 
speed. The shock thus produced may result in injury to the impeller, 
shaft coupling or driving motor. 



Counter Shaft- 



171. A Centrifugal Pump Should Be Started With The 
Discharge Valve Closed. — This is generally necessary (Sees. 
168 to 170) to facilitate priming 
of the pump. But aside from 
this consideration, closure of the 
discharge valve while the pump 
is being started is advisable in 
order that the full load may be 
imposed gradually (Sec. 146) 
upon the driving motor or 
mechanism. When running 
against a closed discharge valve, 
a centrifugal pump requires only 
from 35 to 50 per cent, of the 
power which it consumes when 
running at its economic dis- 
charge capacity. 




Note. — A Centrifugal Pump Can 
Be Run With The Discharge Valve 
Closed Without Generating An 
Excessive Pressure and, there- 
fore, without danger of rupturing 
the casing. But it should not be per- 
mitted to run in this manner longer than about 20 min. at a time. 
When running against the closed discharge valve the propeller churns 
the water in the casing. Churning of the water develops heat therein. 
If continued long enough it may result in the water becoming heated 



Fig. 173. — Centrifugal Pump 
Equipped With Branch Discharge Pipe 
To Prevent Churning When Pump Runs 
Against Closed Discharge Valve. 



150 



STEAM POWER PLANT AUXILIARIES 



[Div. 4 



Thrust Bear/na-. ^^ fft & 



'--Discharge. 



to a very high temperature. This might be dangerous, due to the 
tendency of the rotating parts to expand until seized by the bearings. 

Where a centrifugal pump is driven from a line-shaft, in a factory or 
mill, it may be inconvenient to shut down during lulls in the demand for 
a delivery of water from the pump. But in such cases a small branch 
discharge pipe (Fig. 173) should be run from the discharge nozzle to the 
suction well, so that a small quantity of water may pass through the 

pump and thus prevent churning and 
heating when the discharge valve is 
closed. Where an independent driv- 
ing apparatus, as an electric motor 
(Fig. 174) is used, it should be shut 
down when a delivery of water is not 
desired. 

172. To Start A Centrifugal 
Pump the discharge valve should 
first be closed and the pump 
primed. First turn the impeller 
over a few times by hand to allow 
all air to escape by way of the 
air cock at the top of the casing. 
After the pump has been fully 
primed (Sees. 168 to 170) it may 
be started. The priming con- 
nection should be closed as soon 
as the impeller shaft begins to 
turn. The discharge valve 

Fig. 174.— Submersible Type Of Vertical should remain closed Until full 
Centrifugal Pump Installed In Sump. gp ee d is attained. It should 

then be opened slowly so that the motor may pick up the full 
load gradually. If the pressure does not build up as the speed 
increases, the pump is not thoroughly primed. In this event the 
pump should be stopped and reprimed. Finally, the bearings 
should be examined to see that the automatic oilers are working 
properly, and the packing glands should be adjusted to allow a 
reasonable leakage from the stuffing-boxes. Leakage from a 
stuffing-box indicates that water is being supplied to the 
water-seal ring which is placed in the stuffing-box and which 
prevents air from being drawn into the casing. Usually, the 
nuts on the gland bolts can be drawn sufficiently tight with the 
fingers. 




s^Sfeis 



' Piimp' 



'■"-Strainer' 



Sec. 173] CENTRIFUGAL AND ROTARY PUMPS 151 

Note. — Under No Circumstances Should A Centrifugal Pump 
Be Run In The Wrong Direction. The right direction is generally 
indicated by an arrow (Figs. 154 and 155) cast upon the casing. When 
a centrifugal pump is run in the wrong direction, certain interior parts, 
which are held in place by screw threads, are liable to unscrew and 
thereby wreck the pump. 

173. Electrically-Driven Centrifugal Pumps Should Be 
Operated Under A Steady Voltage. — No attempt should be 
made to operate an ordinary motor-driven centrifugal pump 
with electric current from any circuit, as a street-railway 
trolley circuit, the voltage of which fluctuates widely. 

Note. — If A Motor-Driven Centrifugal Pump Is Designed To 
Run At A Speed Corresponding To The Motor Speed at the maxi- 
mum value of a fluctuating voltage, it will deliver little or no water when 
the voltage is low. On the other hand, if it is designed to give the desired 
capacity with the motor speed corresponding to the minimum value of 
the voltage, the motor may be seriously overloaded when the voltage 
rises to its maximum value. 

174. Centrifugal Pumps Are Not Difficult To Maintain 

in serviceable condition. This is due chiefly to the absence 
of reciprocating parts in their structural details. The prin- 
cipal details to be looked after are the shaft-bearings, stuffing- 
boxes, and wearing-rings. 

Note. — Before A New Centrifugal Pump Is Put In Service 
the bearings should be carefully cleaned with kerosene or gasoline. The 
oil-wells should then be filled with a good quality of mineral oil, such as 
is especially prepared for motor bearings. The oil should be strained to 
insure that no gritty matter is mixed with it. The oil in the wells should 
be changed at proper intervals, perhaps every two weeks in the majority 
of cases. At such times the bearings should be thoroughly washed with 
kerosene. 

Note.— When A Centrifugal Pump Is Used For Moving Corro- 
sive Liquids or sewage, the water used in water-seal connections of the 
stuffing-boxes should be piped from some clear-water source. In such 
cases the now unnecessary openings which would otherwise be in the 
casings should be plugged. 

Note. — The Stuffing-Boxes Of Centrifugal Pumps Should Be 
Packed with loose braided cotton packing impregnated with graphite. 
Ordinary flax packing should not be used, inasmuch as the friction be- 
tween this kind of packing and a rotating rod is apt to be excessive. 

Note. — It Is Generally Advisable To Drain The Casing Of A 
Centrifugal Pump When The Pump Is Out Of Service. This is 



152 



STEAM POWER PLANT AUXILIARIES 



[Div. 4 



imperative where the pump is exposed to freezing temperature. The 
pump may be drained by removing the plug (Fig. 131) in the bottom of 
the casing. 

Note. — Where A Vertical Centrifugal Pump Is Required To 
Operate In A Submerged Position (Fig. 174) the shaft connection to 
the motor should be so designed as to remove the ball thrust-bearing, 
which sustains the weight of the shaft and impeller, entirely from contact 
with the liquid that is being pumped. Adequate lubrication of the bear- 
ing cannot otherwise be secured. 

175. A Rotary Pump (Fig. 175) is a positive-action dis- 
placement pump. It should not be confused with the centri- 
fugal pump. While the 
motion of both types of 
pumps is one of rotation, 
the principles involved are 
entirely different. 

Discharge Noiz/ej. 



Discharge. 
Nozz/e- 




Suction 
Nozz/e ^ 



Fig. 175. — Positive Action Rotary 
Pump. (Goulds Mfg. Co.) 



-Suction 
No. ' 



Fig. 176. — Showing Operation Of A Rotary 
Pump. 



176. The Action Of The Rotary Pump may be understood 
by a consideration of Fig. 176. Suppose the pump is fully 
primed, that is, the casing and suction pipe are completely 
filled with water. The shafts, S, are driven in the direction 
as shown, by a pair of spur gears which are outside of the 
casing, C, and are not shown in the illustration. The liquid 
is engaged by the teeth of the lobe gears, G, and being thus 
confined in the spaces, B, by the lobe gear teeth and the casing, 
is carried upward by the rotation of the gears to the discharge 
outlet, D. The teeth of the lobe gears are so designed that at 
every instant they mesh with each other, thus preventing the 
water from returning to the suction side between the gears. 



Sec. 177] CENTRIFUGAL AND ROTARY PUMPS 153 

177. The Advantages Of The Rotary Pump are: (1) Low 
first cost. (2) Small dimensions. (3) Ease with which it may 
be cleaned. 

178. The Disadvantages Of The Rotary Pump are: (1) Noisy 
in operation. (2) Inefficiency due to the slip which is caused by 
the wear on the lobe gear teeth. (3) Low speed, which usually 
necessitates the use of reduction gears if motor or steam 
turbine drive is employed. 

179. The Services For Which Rotary Pumps Are Most 
Commonly Used are: (1) Fire protection. (2) Pumping of 
oils, chemicals, cider, vinegar, etc. (3) Circulating cooling- 
water for gas engines. (4) Circulating oil for pipe-cutting and 
threading machines. (5) Factories in which food products are 
handled in liquid form. The feature which adapts the rotary 
pump to most of these services is the ease with which it can be 
cleaned. These pumps are manufactured in sizes ranging 
from the small hand-operated size to those having a capacity 
of 900 to 1,000 gal. per min. against a 230-ft. elevation. 

QUESTIONS ON DIVISION 4 

1. Why was the development of the centrifugal pump retarded until recently? 

2. What is a centrifugal pump? 

3. Explain, using a sketch, the theory of the centrifugal pump. 

4. Upon what law of physics is the peripheral speed of a centrifugal-pump impeller 
based? 

5. What is the total head pumped against? 

6. Upon what factors depend the quantity of water which a centrifugal pump will 
deliver? 

7. What theoretical relations exist between the speed of the impeller and the quantity 
of water delivered? The head produced? The required power? 

8. What can be said concerning the applications of the turbine and volute pumps? 

9. How does increasing the number of stages increase the head produced? 

10. What are the forces which tend to unbalance the impeller? 

11. Explain two methods of counteracting these forces. 

12. How is end-thrust taken care of in double-suction pumps. 

13. Name some advantages and disadvantages of the open impeller. Of the enclosed 
impeller. 

14. What is the chief disadvantage of a vertical-shaft centrifugal pump? 

15. In what ways are centrifugal pumps classified? 

16. What are the characteristics of a centrifugal pump? How are they obtained? 
Draw and explain a characteristic graph for a centrifugal pump. 

17. Give four methods of driving centrifugal pumps and tell wherein each method 
is applicable. 

18. Why is a flexible coupling used to direct-connect a centrifugal pump to its motive 
power? 

19. Name the more common services to which centrifugal pumps are applicable. 

20. Why must a centrifugal pump be installed with its center-line below the supply- 
water level when handling hot water? 



154 STEAM POWER PLANT AUXILIARIES [Div. 4 

21. Upon what two factors in the installation of a centrifugal pump does successful 
operation of the pump mainly depend? 

22. What is the highest suction lift generally advisable for centrifugal pumps? 

23. What is a right-hand centrifugal pump? A left-hand centrifugal pump? 

24. What consideration, in any case, should decide whether a right-hand or a left- 
hand centrifugal pump should be installed? 

25. Explain how a centrifugal pump should be leveled and grouted. 

26. Under what circumstances would it be advisable to make the suction pipe of a 
centrifugal pump two or more sizes larger than the suction nozzle? 

27. What should be the least depth of submergence of a vertical suction pipe? Why? 

28. How should the suction line of a centrifugal pump be valved? 

29. Why is it inadvisable to draw water from a battery of driven wells with but one 
centrifugal pump? 

30. In centrifugal-pump operation, how may trouble due to air-impregnated suction- 
water be avoided? 

31. What is the principal consideration in determining the proper size of discharge 
piping for a centrifugal pump? 

32. What is the function of a check-valve in the discharge line of a centrifugal pump? 
Of a gate valve in the discharge line? 

33. What is meant by priming a centrifugal pump? 

34. Why is it detrimental to run a centrifugal pump without liquid in the casing? 

35. Explain the operation of priming a centrifugal pump with an ejector where no 
foot-valve is used. Where a foot-valve is used. 

36. Why is it detrimental to prime a centrifugal pump while the impeller is in motion? 

37. Why should a centrifugal pump be started with the discharge valve closed? 

38. Why is it generally objectionable to run a centrifugal pump continuously with 
the discharge valve closed? 

39. How may a centrifugal pump be piped so that it may, with safety, be run con- 
tinuously with the discharge valve closed? 

40. Explain the procedure of starting a centrifugal pump. 

41. Why is leakage from the stuffing-boxes of a centrifugal pump permissible? 

42. What damage may result from running a centrifugal pump in the wrong direction? 

43. Why is it inadvisable to operate a motor-driven centrifugaL pump with electric 
current from a trolley circuit? 

44. Mention some precautions which should be adopted regarding the lubrication 
of centrifugal-pump bearings. 

45. How may clear water be obtained for sealing the stuffing-boxes of a centrifugal 
pump if the pump is handling sewerage? 

46. What kind of packing should be used in the stuffing-boxes of a centrifugal pump? 

47. Explain the action of a rotary-pump. 

48. Name its advantages. Its disadvantages. 

49. To what services is it applicable? 

PROBLEMS ON DIVISION 4 

1. What must be the theoretical peripheral velocity of the impeller of a centrifugal 
pump to deliver water against a total head of 160 ft.? 

2. If the impeller in Prob. 1 is to be driven at 1,710 r.p.m., what should be its diameter? 

3. A centrifugal pump running at 1,140 r.p.m. produces a head of 90 ft. Assuming 
no head to be lost in friction, what head will the pump produce at 1,600 r.p.m.? 

4. If a centrifugal pump delivers 400 gal. per min. when running at 1,450 r.p.m., 
what will be its capacity when driven at 1,600 r.p.m.? 

5. A belt-driven centrifugal pump requires 10 h.p. to drive it at 900 r.p.m. What 
should be the width of the belt if the driven pulley is 7 in. in dia.? 



DIVISION 5 

INJECTORS 

180. An Injector (Fig. 177) is a boiler-feeding device con- 
sisting of a group of nozzles so arranged that a jet of steam 
expanding therein strikes a mass of water and condenses. 
Thereby it imparts its velocity and heat energy to the feed- 
water which gains, in this way, sufficient momentum to force 
itself into the boiler against a pressure higher than that of the 
original steam. 



Regulating 
.-Valve 



*" -Steam Supply Line 
Injector-.^ 




■Funnel 







W-SMWRER&W&Q&m 



mm?m$ 



Fig. 177.— Illustrating The Principle Of The Injector. 



Note. — Ordinary injectors can discharge against a pressure greater 
than 130 per cent, of the steam-supply pressure. Special injectors are 
obtainable which will utilize exhaust steam at atmospheric pressure and 
therewith pump water into a boiler containing steam at 80 lb. per sq. in. 
A brief explanation of the principles involved will clarify this seeming 
mystery. 

155 



156 STEAM POWER PLANT AUXILIARIES [Div. 5 

181. The Theory Of The Injector may be explained thus: 
A pound of steam is a reservoir of considerable energy. Ex- 
panding, in a well-designed nozzle, from 150 lb. per sq. in. 
(gage) down to a 24 in. vacuum, 20 per cent, or about Y^ of 
its heat content is changed into mechanical energy of motion, 
or kinetic energy, amounting to 188,000 foot-pounds. If all 
of this kinetic energy could be utilized, it would force 500 lb. 
of water back into the boiler. ' Over 97 per cent, of it, however, 
is changed back again into heat when the steam jet, travelling 
at the rate of 40 mi. per min., projects itself against the 
slowly moving mass of water. 

Note. — The impact of two bodies always results in the generation of 
heat at the expense of kinetic energy. Now, the remaining 3 per cent, 
of kinetic energy, after the 97 per cent, has been reconverted into heat, 
is sufficient, theoretically, to force 15 lb. of water back into the boiler. 
But pipe friction and other losses cut this down to about 13 lb. of water 
pumped per pound of steam consumed at 150 lb. per sq. in. pressure. 
The remaining 97 per cent, of the energy which was changed back into 
heat and the % of the original heat content of the steam, are not lost 
but are absorbed by the feed water and returned to the boiler. 

182. The Essential Parts Of An Injector are shown in Figs. 
177 and 178. These are purposely drawn out of proportion 

so that the characteristic 
shapes of the nozzles can 
be discerned more clearly. 

^ == j ; i^py^== =Q "., " * ' I Explanation. — The steam 

JFJ^ ffi* * " '-'fr ' j c -^ |: - -'" ^-^^ j nozzle, S, (Fig. 178) at the left, is 

" t -- g ^ j*{?^ T "\ ^ , . "" so designed that the steam, in 

iPll Co mh^' n9 ''' : Delivery^ r \ passing through it, loses pressure 
Wafer fi I lube / Tube D T f> ', -, • ° ' , •. 

inlet- [■- 1 :' e , Boiler-- and gams a tremendous velocity. 

1 1 I' II -Overflow & J 

llwai Chamber. When a pound of steam expands 

Fig. 17 8. -Sectional View Of Elementary from boiler pressure to a partial 
Injector. vacuum and to the correspond- 

ing lower temperature, it liber- 
ates heat which is converted into kinetic energy and thereby causes the 
steam to attain a very high velocity. For an explanation of the conver- 
sion of heat energy into kinetic energy, due to expansion through a 
nozzle, see the author's Steam Turbines. The combining tube, C, 
is a cone-shaped nozzle in which the swiftly moving steam jet strikes 
the water and is condensed. The delivery tube, D, is a diverging 
nozzle. It receives the combined jet of water-and-condensed-steam 



Suction 




Overflow Outlet 


Chamber, 


Steam 


Hj_nt 


Steam 


,-Nozile 


■ V |U--'| Valve 


Inlet-. | 


\ Overflow-. 





Sec. 183] 



INJECTORS 



157 



and gradually converts most of the kinetic or velocity energy of the 
jet into static energy or pressure. This is needed to overcome the 
head against which the injector is discharging. Overflows, H, are 
slots or spill-holes, usually located in the combining tube, to permit 
excess water or steam to escape when starting up. The waste- 
valve, V, may be a stop valve 



BfJEDTJCOTl 




but is usually a lift or swing 
check which closes automatic- 
ally in case that a partial 
vacuum is formed in the over- 
flow chamber, O. Thus, V, 
prevents the inrush of outside 
air that would tend to scatter 
the jet. The water in the suc- 
tion chamber, W, is drawn into 
the combining tube by the par- 
tial vacuum which is due to the 
continuous condensation of the 
steam therein. 

183. Injectors Are 
Classified as: (1) Lifting. 
(2) Non-Lifting; depend- 
ing on whether or not a 
partial vacuum is created 
in the suction pipe when 
starting up. A non-lifting 
injector must always be 
placed below its source of 
feed water on this account. 
Injectors that have one set 
of nozzles (L Fig. 179) for 
lifting the water and an- 
other set, F, for forcing it 
into the boiler are called 
double-tube injectors. Those 
which accomplish the same 
result with only one set of 

nozzles (Fig. 180) are called single-tube injectors. If the oper- 
ation of an injector automatically re-establishes itself after an 
interruption in steam or water supply, it is said to here-starting, 
or, more usually, automatic. But when the injector must be 



Fig. 179. — The "Hancock Inspirator," A 
Double-Tube Injector. v It is claimed that it 
will, without adjustment, operate on steam 
pressures ranging from 15 to 240 lb., lift water 
25 ft. or take it under head. Lift 140-deg. water 
3 to 4 ft., lift 90-deg. water 25 ft., and with 
45 lb. steam presssure will lift water 25 ft. 
and elevate it 112 ft. above inspirator.) 



158 



STEAM POWER PLANT AUXILIARIES 



[Div. 5 



Steam- 



Steam Nozzle 



manually re-started, before it will continue to operate, it is 
said to be positive. Automatic adjustment for variations in 
steam pressure or in height of lift and temperature of feed 
water is a feature of self adjusting injectors. All double-tube 
injectors, and a special type of single-tube injector which has 

a moving combining tube, 
belong to this class. The 
Sellers self-acting injector 
is both self adjusting and 
re-starting. 

184. How An Automatic 
Injector Works is indicated 
by Fig. 180 which shows a 
section through a Pen- 
berthy Automatic In- 
jector. This is a single 
tube, re-starting, lifting- 
type injector: 

Explanation. — Steam enters 
at the top, and, expanding in the 
steam nozzle, R, rushes through 
the draft-tube, S, carrying with 
it enough entrained air to create 
a partial vacuum in suction 
chamber, B. Unable to dis- 
charge against the boiler pres- 
sure, this steam escapes through 
the large opening above the 
sliding washer, T, and through the overflow opening, D, via P and O 
to the atmosphere. The partial vacuum in B has already lifted water 
into it, and this water has condensed part of the steam. As more and 
more of the steam condenses, the jet becomes more compact and finally 
becomes sufficiently small to pass through the least diameter of the 
combining tube, C. Thence it passes through the delivery tube, Y, 
and a check valve (Fig. 187) to the boiler. 

The swiftly-moving jet of water-and-condensed-steam creates a par- 
tial vacuum in tube C. This draws the loose washer, T, up against its 
seat. Thereby is prevented any inrush of air which would scatter the 
jet. The closing of T also prevents any loss of feed water through it. 
If the steam or water supply becomes interrupted, the jet is destroyed 
and the vacuum above T is lost. This allows T to drop down to its 
original position. Hence, upon the resumption of the steam or water 
supply, the operation just described is repeated. 




v -7b 
Boiler 



Fig. 180. — "Penberthy" Single-Tube Auto- 
matic Injector. 



Sec. 185] 



INJECTORS 



159 



185. How A Positive Injector Works is indicated in Fig. 181 
which shows a section view through a Metropolitan Model 
Injector. This is of the positive, double-tube, lifting type and 
is operated entirely by one handle. 

Explanation. — When handle, //, is pulled back slightly, steam is 
admitted to the lower lifting nozzle, N, through the opening of the 
auxiliary valve, A, and of the regulating value, R. The lifting nozzles. 
iVand C, now begin to operate. The excess steam escapes through the 
intermediate overflow valve, 0, and thence to the atmosphere, through 
the final overflow or waste valve F. As soon as water is lifted, it will 
reach the overflow through C. The operator then pulls the handle back 



hancf/e 



.--Bell Crank 




Fig. 181. 



■Overflow 
'-Intermediate Overflow Valve- 

-Suct'on 

Metropolitan" Model-O, Double-Tube Injector. 



Mixing; Nozzle--'' 
Suction Chamber-- 



gradually, admitting steam into the main nozzle, M, through the steam 
valve, V. The action of this steam in passing through the remaining 
nozzles has already been explained. By the time the handle has been 
pulled back as far as it will go, the injector is feeding into the boiler 
through check valve, D, and the link, L, has moved to the left far enough 
to close the waste valve, F, by means of the bell crank B and the stem S. 
Regulating valve R is used to control the supply of water to the injector. 

186. The Advantages Of An Injector are: (1) Simplicity. 
(2) Compactness. (3) Low first-cost. (4) High temperature of 
feed-water delivered. (5) Ease of operation. (6) Low cost of 
upkeep and repairs. (7) High thermal efficiency, about 99 per 
cent, of energy put into it is utilized. The absence of any 
moving parts is responsible for most of these advantages. 



160 STEAM POWER PLANT AUXILIARIES [Div. 5 

There are practically no packing glands to be renewed and no 
parts to be lubricated. 

Note. — Cold Feed- Water Sets Up Strains That Endanger The 
Structural Strength Of A Boiler. Hence, an injector is of peculiar 
advantage on locomotives where the lack of space and the use of the 
exhaust steam for stack draft prevent the installation of feed-water 
heaters. These same two conditions render the injector peculiarly 
applicable on locomotives because of its compactness and because it is 
many times more economical than the feed pump if the exhaust from 
the latter is not used to heat the feed-water. Used as an emergency feed, 
an injector involves a minimum of overhead expense. 

187. The Disadvantages Of The Injector are: (1) Inability to 
handle water which is very hot. (2) Irregularity of operation 
under extreme variation in steam pressure, in temperature of 
inlet water and in quantity of water handled. (3) Efficiency 
as a pumping unit is extremely low, never over 1 or 2 per cent; 
that is, when used in ordinary pumping service — not for 
boiler feeding — an injector does not compare at all favorably 
with ordinary pumps in economy. Few injectors can handle 
water at 150°F. and most of them become inoperative at 
much lower inlet water temperatures. This is the real reason 
why injectors are not extensively used in large power plants. 
Such plants always have an ample supply of exhaust steam, 
available from the auxiliaries. If this steam is not used to 
heat the feed-water it will be wasted. 

Note. — A Feed- Water Heater Placed On The Suction Side Of 
An Injector Would Heat The Water Too Hot For Its Successful 
Operation. Placed on the discharge side, a feed-water heater would be 
inefficient because the injector would deliver water to it at such a high 
temperature that the heater would not abstract much additional heat 
from the exhaust steam. To heat feed-water with live steam, when 
exhaust steam is available, results in poor economy. The irregularity 
of operation due to variations mentioned above is not, in situations for 
which the injector is adapted, a serious drawback and necessitates only 
a reasonable amount of attention from the operator. 

188. The Applications Of Injectors Of The Different Types 

will now be considered : Whenever it is necessary or desirable 
to locate the injector above the source of feed, the lifting type 
must be used. This is especially true in locomotive practice 
where it is very advantageous to have the injector where the 



Sec. 189] INJECTORS 161 

engineer can see the overflow outlet. The non-lifting type is 
simpler, cheaper and of special advantage where scale-forming 
feed-water is used, because it will not clog up readily and is 
very easy to clean. Double-tube injectors will handle hotter 
feed-water through higher lifts than will those of the single- 
tube type. Hence they are used exclusively on locomotives 
as a main feeding device, and, extensively, on board ship and 
in stationary power plants for emergency boiler-feeding. Re- 
starting injectors are used on small boats, traction and logging 
engines, and in small power plants. They are of special ad- 
vantage for boats, road engines and similar applications 
because the sudden interruption of water supply, due to jar or 
to movement of the boat, will be taken care of by the " auto- 
matic" feature. The "self acting" injector was designed for 
locomotive use but is applicable where any double-tube type 
is necessary. Injectors are often used for testing and washing 
boilers, feeding compound into boiler, and similar services. 
189. A Simple Approximate Equation Of The Injector, which 
shows the relation between : pounds of water pumped per pound 
of steam, the initial temperature of the steam, and the final tem- 
perature of the condensed steam is given below. It is similar to 
one proposed by Julian Smallwood in his Mechanical Labora- 
tory Methods. In this equation radiation losses and the 
amount of heat which is changed into work are neglected. 
These two quantities amount to only lj^ per cent, of the 
total heat energy involved. See derivation below. 

(62) W sw = xHv t {Tfs ™ Tfd) (lb. water /lb. steam) 

1 fd ~ i- fi 

Wherein (see Fig. 182) W sw = pounds of water pumped per 
pound of steam, x = quality or dryness of steam, expressed 
decimally; if steam contains 1 per cent, of moisture, then x = 
0.99; a working average value for per cent, moisture is 2 per 
cent., in which case x = 0.98. H v = latent heat of vaporiza- 
tion of steam at the absolute pressure, P a , at which the injector 
is receiving steam, as taken from a steam table, in B .t.u. T/ s = 
temperature of the steam, at absolute pressure P a , in degrees 
fahrenheit. T fd = final temperature of condensed steam = 
temperature of feed water discharged into boiler, in degrees 
li 



162 STEAM POWER PLANT AUXILIARIES [Div. 5 

fahrenheit. T/% = temperature of intake water to injector, in 
degrees fahrenheit. 

Note. — The Measure Of The Economy Of An Injector is the 
weight of water pumped per pound of steam used. This value may be de- 
termined by applying For. (62). 

Derivation. — When 1 lb. of steam at some absolute pressure P a lb. 
per sq. in., is condensed and then cooled down to a temperature of T f d 
deg. fahr., it gives up a quantity of heat = B.t.u. = xH v + (T/ s — Tfd). 
Now, each 1 lb. of water pumped into the boiler absorbs heat energy 
= B.t.u. = Tfd — Tfi. Then, neglecting the radiation losses and the 
amount of heat which is changed into work (both of which amount to 
only 13^ per cent, of the total heat energy involved), the following ap- 
proximate relation exists in the injector, because the heat absorbed by 
the water must just equal the heat given up by the steam: 

(63) Heat absorbed by water pumped = Heat given up by steam used. 

(64) Heat absorbed per 1 lb. of water pumped = T fd — T fi 

(64 A) Heat given up per 1 lb. of steam used = xH v + (T fw — T fd ) 

Then, if W w = weight of water pumped, in pounds, and W s = weight of 
steam used, in pounds, it follows from For. (63) that: 

(65) W w (T fd - T fi ) = W s [xH v + (T fs - T fd )\ 
Now, transposing: 

(66) W7 = T f <-T, — . 

But if W sw is taken to represent pounds of water pumped per pound of 
steam used, then: W sw = W TO /W S . Now substituting this W sw for its 
equivalent in For. (66), there results For. (62): 

(67) W sw = xHv + {Tfs ~ Tfd) (lb. water /lb. steam) 

lfd — i/i 

Note. — To determine the value of W sw for any injector, it is (assum- 
ing that the quality of the supply steam is known, Author's Practical 
Heat, Div. 19) merely necessary to observe (Fig. 182) the intake and the 
discharge-water temperatures at the injector, observe the steam pres- 
sure, substitute in For. (62) and solve. 

Example. — In testing an injector (Fig. 182) the inlet-water tem- 
perature was 63 deg. fahr., the discharge-water temperature was 202 
deg. fahr., and the steam pressure, as indicated by the gage, was 105 lb. 
per sq. in. The moisture content in the steam was 2 per cent. How 
many pounds of water was this injector pumping per pound of steam 
which it used? 

Solution. — The quality of the steam = x = 1.00 — 0.02 = 0.98. 
The latent heat of evaporation of steam, as taken from a steam table- 



Sec. 190] 



INJECTORS 



163 



at 105 lb. per sq. in. gage (= 105 + 14.7 = 119.7 lb. per sg. in. ab- 
solute) is 877 B.t.u. The temperature of steam at 119.7 lb. per sq. in. 
absolute, as taken from a steam table, is 341 deg. fahr. Now substitute 
in For. (62): W sw = [xH v + (T fs - T fd )]/(T fd - T fi ) = [0.98 X 877 + 
(341 - 202)] -=- (202 - 63) = [859.5 + 139] -=- (139) = 998.5 + 139 = 
7.18 lb. of water per lb. of steam. 



i obobo l ob Wfc) Supply-- 




Fig. 1S2. — Injector Arranged For Testing. 



190. To Compute The Horsepower Actually Delivered By 
An Injector, apply For. (24). The amount of water which the 
injector is handling, may be determined by weighing the water 
before it is pumped. 

191. The Performance Of An Injector Is Influenced By 
The Following Important Factors: (1) Temperature of inlet 
water. (2) Height of suction lift. (3) Steam pressure. The 
action of an injector depends upon the condensation of the 
steam jet by the incoming water. If this water is too warm, 
the injector will not start. This limit is called the over- 
flowing temperature. After the injector has started, it is 
possible to operate with an intake water of a higher tempera- 



164 



STEAM POWER PLANT AUXILIARIES 



[Div. 5 



ture, up to a certain limit called the breaking or limiting 
temperature. 

Note. — Fig. 183 shows how these two temperatures vary with the 
steam pressure. Fig. 184 shows how variations in the feed-water tem- 




25 50 75 100 125 150 115 200 225 250 215 300 325 
Steam Pressure in Lb. per So). Inch 

Fig. 183. — Limiting And Overflowing Temperatures. (This figure was taken 
from page 135 of Kneass' Practice And Theory op the Injector.) 

peratures affect delivery temperature of feed- water, capacity of injector 
and pounds of water pumped per pound of steam. The height of suc- 
tion lift affects the capacity of an injector as shown in Fig. 185 taken 
from 8 tests of a "Penberthy" Size D Automatic Injector operating at 
80 lb. per sq. in. steam pressure and taking feed-water at 74 deg. fahr. 

220 280 c 14 




Inlet Water Temperature °F 
Fig. 184.— Test Results Of A "Sellers" No. 8 Self-Adjusting Injector. 

Fig. 186 shows the variation in capacity of the same Penberthy Injector 
operating under different steam pressure but with the height of lift and 
inlet water temperature constant at 4 feet and 74 deg. respectively. 

Note. — The Reason Why The Water Pumped Per Pound Of 
Steam Decreases With An Increase In Steam Pressure (Fig. 184) 



Sec. 192] 



INJECTORS 



165 



is that the mechanical work done by the injector, in pumping a given 
weight of water into the boiler, increases almost in proportion to the 
steam pressure while the heat content of the steam, and therefore its 
ability to do work, increases but slightly. Between 100 to 200 lb. per 
sq. in. pressure, the heat content of the steam increases by less than 1 
per cent. 




•G450 

c 

d-400 

c 

o35Q 



-Test Results Of A 
Penberthy Automatic Injector 
— 1 — T— (Size D.) — 1 — I — 1~ _ 

Steam Pressurz^gC 1 lb Rvofwater At 7-TF. 



'0 2 4 6 8 10, 12 14 16 18 Z0 
Height Of Lift In Feet 



Fig. 185.— Graph Of Test Results For 
A "Penberthy" Size D Automatic In- 
jector Showing Relation Between Ca- 
pacity And Height Of Lift. 



L..1-., ., 


7S n 2 -% 


= H V 


I - 3 £ 


4 \ 


£ t -^£r 




I 700 - (SzeD) 












."t 




SX 


5600- _ 



Z0 40" 60 '"80 100 120 140160 
Steam Pressure Lb. Per Soj.ln. 

Fig. 186. — Graph Of Test Results 
Of A "Penberthy" Automatic In- 
jector Showing Relation Between 
Steam Pressure And Capacity. 



192. The Selection Of An Injector requires a careful con- 
sideration of the three factors discussed in Sec. 191. Select 
an injector with a capacity in gallons per hour that is 30 per 
cent, in excess of the amount of water normally used. If the 
amount of water evaporated per hour is not known, approxi- 
mate values computed from the following equations, taken 
from Sellers' Restarting Injector, may be used : 

For horizontal or vertical tubular boilers: 

A bh 



(68) Gal. per hr. = 
For water tube boilers: 

(69) Gal. per hr. = 
For flue boilers: 

(70) Gal. per hr. = 



3.2 



2.42 



A-bh 
1.17 



Wherein Abh = area of boiler heating surface, in square feet. 

193. The Question Of What Type Of Injector To Use For 
Any Given Service has been previously discussed in Sec. 
188. It is always best to inform the manufacturer as to the 
height of lift and average temperature of feed water and the 



166 



STEAM POWER PLANT AUXILIARIES 



[Div. 5 



maximum, minimum and average steam pressure, as well as 
the required capacity of the injector. The injector will not 
operate at more than its maximum or less than its minimum 
capacity. Table 194 shows the list prices and other data 
for automatic injectors of a well-known make. 

194. Table Showing Capacities, Pipe Connections And 
Approximate Weight Of Injectors. 









Capacity gal. per hr., 




Manu- 


Approxi- 


Pipe 


1 to 3 ft. lift, 60 to 110 


Shipping 


facturer's 


mate 


connec- 


lb. per sq. in., 


weight, 


size, 


price, 


tion, 


steam pressure 


boxed, 


designation 


dollars 


in. 




lb. 












Maximum 


Minimum 







15.00 


H 


60 


35 


2.4 


00 


16.00 


% 


80 


45 


2.5 


A 


18.00 


y* 


135 


70 


3.5 


AA 


20.00 


X A 


180 


100 


3.5 


B 


25.00 


% 


260 


140 


5.5 


BB 


30.00 


H 


360 


180 


5.5 


C 


40.00 


l 


475 


250 


8.0 


CC 


45.00 


l 


600 


325 


8.0 


D 


55.00 


1H 


800 


425 


12.0 


DD 


60.00 


m 


1,000 


525 


12.0 


E 


75.00 


IH 


1,400 


740 


25.0 


EE 


90.00 


IH 


1,900 


850 


25.0 


F 


110.00 


2 


2,400 


1,275 


37.9 


FF 


125 . 00 


2 


3,000 


1,600 


39.0 


G 


150.00 


2H 


3,600 


1,875 


75 


GG 


200 . 00 


2M 


4,200 


2,150 


75.0 



195. In Installating Injectors the typical piping scheme 
shown in Fig. 187 may be followed. The size of pipe to use 
for injectors of one make can be found in Table 194 under 
Pipe Connections. The steam, suction and discharge pipes 
are all of the same size except that, in the case of a suction lift 
exceeding 10 ft. or of a long length of suction line, a pipe one 
or two sizes larger should be used therefor. 

Explanation. — In Fig. 187, the steam line should be tapped into the 
highest part of the boiler and lagged all the way to the injector, if possible. 



Sec. 195] 



INJECTORS 



167 



C is a globe-valve. The discharge line should follow as near a straight 
line as possible to the boiler-feed inlet and should be securely fastened 




P§s> 






ri//j to Prevent Splashing 



Fig. 187. — Piping Of An Injector. 

throughout its entire length. A check-valve, E, must be installed as 

shown. A is a globe stop valve which can be used to cut off the boiler 

pressure from the check-valve so it may be opened for repair. The 

overflow is usually piped as shown. It is, usually, 

best not to discharge the overflow into the hot-well 

or feed supply as it may cause the suction water 

to become too hot to be lifted. The overflow 

line must always be open at G to the atmosphere. 

The funnel, F, may be an ordinary one of sheet 

metal or one of the special "non-splash" types 

(Fig. 188) on the market. 

The suction line must be absolutely air tight and 
as free from elbows and bends as possible. The 
globe angle valve B takes the place of one elbow. 
The strainer, S, should not have any opening in 
it as large as the steam nozzle in the injector and 
should have a combined area of opening several 
times as great as the suction pipe itself. Figs. 182, 
189 and 190 show commercial strainers. The distance, h, should always 
be below 20 feet and much less than that if possible. Injectors are on 




Fig 
Splash ' 



18 8.— " Non- 
Funnel For In- 



jector Overflow Pipe. 



168 



STEAM POWER PLANT AUXILIARIES 



[Div. 5 



the market that will lift 25 feet. But high lifts reduce the capacity of an 
injector as well as its ability to handle hot water. Further they make 





Fig. 189. — Hose-Connection Strainer For Fig. 190. — Pipe-Connection Strainer For 
Injector Suction-Pipe. Injector Suction-Pipe. 



idtpH 



E Overhead Tank HL 

Dcz3 

T 




■— — ii — — Overflow line, 
I li IL • ! 






Fig. 191. — Injector Fed From Overhead Tank. 



starting difficult and operation impossible when there is even a small 
leak in the suction line. 



Sec. 196] INJECTORS 169 

If the injector is fed from an overhead tank (Fig. 191) or from city 
supply under pressure, it is advisable to insert an additional valve, D, 
(Fig. 187) which can be permanently set so as to throttle the pressure 
down to the desired limit. Then, the valve, B, is used only for opening 
and closing the feed line. All injectors should be braced, especially 
those which are operated by handles. After installing the piping, it 
should all be blown out with steam before connecting up the injector. 

196. In Operating Injectors, the procedure is as follows: 
To start an automatic injector be sure that valve A (see Fig. 187) 
has been left open. Open slowly steam valve, C, wide, now 
open suction valve, B, wide. Then throttle it down until 
there is no discharge from the overflow. If the suction valve 
is wide open and steam still escapes from the overflow, it will 
then be necessary to throttle the steam-supply valve. If the 
discharge from the overflow is cool water, then the suction 
valve must be throttled. If there are no unusual changes in 
conditions, the suction valve B can be adjusted to give proper 
supply and then be permitted to so remain. An injector like 
that shown in Fig. 181 is operated entirely by one lever as 
described in Sec. 185. 

197. Injector Troubles And Their Correction are discussed 
below. The more important ones are listed. The correction 
of other difficulties can, usually, be effected through a con- 
sideration of the information given here : 

If An Injector Fails To Lift Water, it may be due to the following 
causes: (1) Leak in suction line. (2) Water too hot. (3) Steam pressure 
too low for the lift. (4) Suction strainer clogged. (5) Wet steam. (6) 
Nozzles of injector clogged up or covered with scale. (7) Waste or overflow 
valve stuck or leaking. (8) End of suction line not below water. (9) 
Suction hose collapsed by partial vacuum. To test for leaks in suction 
line, screw a cap on the end of line in place of the strainer. Then wedge 
the waste valve shut with a piece of wood. When steam is turned on, 
the leaks will be detected easily. Steam is liable to be wet unless taken 
from the top of the boiler and led directly to the injector. If nozzles 
are clogged with scale they can be removed and cleaned. Coatings of 
lime can be removed by soaking the nozzle several hours in a solution of 
ten parts water and one part muriatic acid. 

If An Injector Lifts Water But Does Not Deliver To The 
Boiler the trouble may be due to (1), (3), (5), (6) and (7) of the above 
and may also be caused by: (10) Faulty boiler check valve. (11) Obstruc- 
tion somewhere in delivery pipe. In case of the latter two difficulties, 
close valve A and examine the check valve. If it is lifting properly 



170 STEAM POWER PLANT AUXILIARIES [Div. 5 

leave the cap off and take out the disk. Then start the injector. If a 
full stream of water shoots out of the check valve, then there is an ob- 
struction between it and the boiler (most probably inside at the opening 
of the feed pipe). 

If The Injector Starts But Breaks, the trouble may be due to 
(1), (3), (6), (11), and also to (12) An improper adjustment of the water 
supply. If water at the overflow is hot then the supply is inadequate 
and should be increased by opening valve B wider. If it is cold then the 
supply should be throttled by means of valve B. 

When Steam Appears At The Overflow the fault may be (2) or 
(4) or (13) Too-high steam pressure for the lift. In this case throttle down 
the valve C until the overflow discharge ceases. Every user of injectors 
should preserve a set of directions for removal of injector parts and should 
have available spare nozzles for repairs. Directions are gladly furnished 
by the manufacturers. 

QUESTIONS ON DIVISION 5 

1. Explain how it is that an exhaust steam injector can pump water into a boiler 
against the boiler pressure. 

2. Name four important parts of an injector, giving functions of each. 

3. Distinguish between an automatic and a positive injector. 

4. What is a self adjusting injector and why are all double-tube injectors of this class? 

5. Name and explain six advantages of injectors over feed pumps. 

6. Why are injectors seldom used in large stationary plants? 

7. Why are injectors always used on locomotives? 

8. Explain effect of: (1) Steam pressure. (2) Height of lift. (3) Temperature of inlet 
water upon the capacity of an injector. 

9. Give eight general rules which should be followed in installing injector piping. 

10. Give eight possible causes for an injector's inability to lift water and state the 
correction for each. 

PROBLEMS ON DIVISION 5 

1. The following data were observed during an injector test: Temperature of inlet 
water, 60 deg. fahr. Temperature of discharge water, 200 deg. fahr. Steam pressure, 
100 lb. per sq. in., gage. Moisture in steam, 2% per cent. Find value of W» w or 
pounds of water pumped per pound of steam? 

2. Assume all data in Prob. 1 except temperature of discharge water. Find what 
this temperature will be if W sw = 10? 

3. A water-tube boiler has a heating surface of 500 square feet. What size injector, 
as given in Table 194, should be used to handle the feed water? 

4. What size steam, suction, and delivery pipes should be used in Prob. 3 if the 
height of lift is 8 ft.? If it is 15 ft.? If it is 20 ft.? 



DIVISION 6 



BOILER-FEEDING APPARATUS 



198. Apparatus For Feeding Water To Steam Boilers 

includes devices of three principal types: (1) Injectors, 

(2) Pumps, (3) Return traps or gravity apparatus. Injectors 

for boiler-feed service in stationary power plants are usually 

installed only as stand-by or emergency equipment. Under 

certain conditions, however, they may show an economic 

advantage over other forms of apparatus. Pumps are the 

most important boiler-feeding devices. Direct-acting steam 

pumps (Div. 2), crank-action pumps variously arranged and 

driven (Div. 3) and centrifugal pumps (Div. 4) are all used for 

boiler-feeding as will be explained herein. A few years ago 

the direct-acting steam 

t> [ p\ 

pump was the most 

widely used variety of 
boiler-feed pump and is 
still a very common va- 
riety. The use of gravity 
boiler-feeding apparatus 
is limited largely to small 
steam heating and non- 
condensing power install- 
ations. 



Connection To Fetch 
Wxter Heater-, 




\\\^\\\\\\\\\\\\\\\^ 



Fig. 192. 



-A Direct-Acting Steam Pump For 
Boiler Feeding. 



Note. — The general rules 
for piping the principal de- 
vices which are used in boiler- 
feeding are similar to those for 
steam piping. (See Div. 11 

for types of joints, specifications and allowable pressures.) There are 
usually at least two feed pumps in a stationary power plant and each 
should be connected to a common header supplying all of the boilers. Ex- 
pansion (see Div. 11) in feed water lines is not as great on the whole as 
in steam lines but must be allowed for nevertheless. The pump inlets 
should be connected so that water may be drawn from two or three sources 

171 



172 STEAM POWER PLANT AUXILIARIES [Div. 6 

(Fig. 192) such as hot-well, feed-water heater and city water-mains so 
that feed water of some sort is always available during repairs or 
emergencies. 

199. When An Injector Is Used As A Pump For Raising 
And Forcing Water And Only As A Pump, it is very inefficient 
inasmuch as it requires about five times as much steam — or 
coal — as does an ordinary simplex or jduplex steam pump to 
do the same work. Hence, as a devifee for merely handling 
water where boilers are not to be fed, the injector is, on an 
economic basis, entirely out of the running. Furthermore, 
there are a number of troubles (Sec. 197) of the injector that 
further limit its usefulness. The injec|or cannot, in practice, 
effectively handle water at temperatures exceeding about 
100 deg. F. This means that it cannot be used advantage- 
ously with water which has been previously heated with the 
feed-water heater. Hence, the injector cannot be used at all 
with an open feed-water heater. It may be used with a closed 
heater installed between the injector and the boiler. 

200. An Injector Will Not Start Whe| Served By A Steam 
Pressure Much Lower Than That For Wnich It Was Designed. 
Assuming that an injector is started ^on the pressure for 
which it was designed, then if the impressed pressure increases 
or decreases materially the injector will cease to work. Nor 
will it start again automatically upon resumption of the 
steam pressure at which it originally started and for which 
the engineer temporarily adjusted it. To again cause it to 
pump water, the engineer must perform anew the starting and 
adjusting process. Furthermore, material change in the 
water pressure of the suction water which is being handled 
by the injector, will cause it to cease operation. This necessi- 
tates a new adjustment and a new start. Often when an- 
injector has been working and has become hot, if for any 
reason it stops or is stopped, it cannot be re-started until it 
has been cooled completely by sousing it with .cold 'water. 
Obviously, all of the above disadvantages restrict the desirable 
applications of the injector for boiler-feed service. Oh the 
other hand, the simplicity, small space occupied, absence of 
moving parts, and low iirst -cost of the injector render its use 
desirable under certain conditions. . - - 



Sec. 201] 



BOILER-FEEDING APPARATUS 



173 



201. The Injector Is Economical For Feeding Boilers In 
A Plant Not Equipped With Means Of Feed-Water Heat- 
ing. — Under this condition the injector (Fig. 193) acts as a com- 
bined pump and pre-heater; and, as such, is almost 100 per 
cent, efficient. The conditions favorable to injector installa- 
tion often obtain in temporary or out-of-the-way plants where 
the equipment must be minimized, and where the saving 
which would occur through the installation of a feed-water 
heater is more than offset by its annual cost; see Sec. 246. Its 




Fig. 193. — An Inspirator Type Of Injector Piped Up For Boiler Feeding. 

feature of pre-heating its feed water, makes the injector addi- 
tionally valuable where cold water is to be fed into the boiler. 
By pre-heating, the strains which cold water would cause in 
the boiler are avoided. The proper combination, however, 
of a pump with a feed-water heater is, as a rule, more satisfac- 
tory than an injector for stationary power plants. Injectors 
are effectively employed on boilers for traction-engines, 
small saw-mill engines, hoisting and logging engines, and on 
locomotives. 

202. The Relative Efficiencies of Steam Pumps and In- 
jectors As Boiler-Feeding Devices are given in Marks' 
Mechanical Engineers' Handbook as follows: The effi- 
ciency of an injector considered merely as a pump is very low, 



174 



STEAM POWER PLANT AUXILIARIES 



[Div. 6 



about 1 to 2 per cent. As a boiler feed pump, in which service 
the heat in the steam consumed is returned (see Sec. 181, 
also Fig. 195) to the boiler, the injector has an individual 
efficiency of nearly 100 per cent. However, the injector is 
not ordinarily the most economical device for feeding a boiler 
since it can handle only cold or moderately-warm water and 
the effect is equivalent to heating the feed water with live 
steam. On the other hand, a pump can handle water which 
has been heated to a relatively-high temperature by exhaust 
steam (from the main or auxiliary engines) which would other- 
wise be wasted. Injector steam consumption is about 400 
lb. per water h.p. hr.; a small direct-acting pump consumes 
100 to 200 lb. 



■moB.t.u. (p^p ' 



&> ^Combined 

5 Frhnrmt-. 




rlnjecfor ^iWRtM..^. y 

.-m.t.u. ~~\^ 

' -86B.t.u. I 1 

Returned As 
Warm Water 

Water60'\ 



Heat Of Steam Above 60'Useol By Engine H60B.f.u. 
Heat Of SteamAbove60'UsedBy Pump 20B.f.u. 

Total - H80B.t.u. 
krReloitjve Heat Consumption - WOO 

Fig. 194. 
Fig. 194. — Direct-Acting Feed Pump, 
the four following illustrations, are B.t. 
engine.) 

Fig. 195. — Injector, No Heater. 




Exhaust-. 



vn^\\Sn?-\ v \V\\ \\\^\ W\vv\\ s\VVnV\\v 



Heat Of Steam Above 60°Useof By Engine 1160 

- " - • loss By Injector 88-86 - Z 

Total - 1162 

B^ Heat Consumption Relative ToA'Ti§r = 0385 

Fig. 195. 

No Heater. (B.t.u. values in this, and 

l. per pound of steam delivered to the 



t/mt.U: 



Part Of 
Exhaust-, 




HeatOf Steam Above60°Useof By Engine IWB.t.u. 
« <• ' » " "Injector 88 " 

" • " . • Returned To Boiler 140 •' 



Net Heat Used 



1108 - 



Fig. 196. — Injector And Exhaust 
Heater. 



Part Of 
Exhaust 




\\\\\\\\\\\\\\\\ 



HeafOf Steam Above 60°Used By Engine IKO&t.u. 
• » » " •■ p um p ft „ 

" " • Returned To 140 • 

Total Heat Used 1040 - 

D-Heat 'Consumption Re/afiveToA'Mo z OJ82 

Fig. 197. — Direct-Acting Feed Pump 
And Exhaust Heater. 



203. Table Showing Relative Economies Of A Non- 
Condensing Plant Using Boiler -Feeding Devices Of Different 
Types. (Based on data by D. C. Jacobus, Kent's Mechani- 



Sec. 203] 



BOILER-FEEDING APPARATUS 



175 



cal Engineers' Pocketbook). See Figs. 194, 195, 196, 197, 
and 198. In each case the values are for the same plant de- 
livering the same power output from its engine. The only 
differences between the cases are in the boiler-feeding and feed- 
water heating arrangements. 







Relative 










steam 


Per cent. 


Refer- 




Equipment 


consump- 


steam 


ence 






tion from 


saving 


letter 






boilers 






u 


Direct-acting steam pump re- 






Fig. 


o3 


ceiving water at 60 deg. fahr. 






194. 


£ 


and forcing it directly into boiler 








13 f- 


at 60 deg. fahr 


1.000 


0.0 


A 


P 1 






fe § 
-B 


Injector receiving water at 60 






Fig. 


o 


deg. fahr., heating it to 146 deg. 






195. 


+= 
^ 


fahr. and forcing it directly into 








boiler at that temperature 


0.985 


1.5 


B 




Injector feeding water through a 






Fig. 




heater in which it is heated from 






196. 


o3 


146 deg. fahr. to 200 deg. fahr. . . 


0.938 


6.2 


C 




Direct-acting steam pump feed- 






Fig. 


t-l 


ing water through a heater in 






197. 


o3 


which it is heated from 60 deg. 








fahr. to 200 deg. fahr 


0.882 


11.8 


D 


13 








Geared nower pump mechani- 






Fig. 




- ' - n g 
ich 
. to 






198. 






0.868 


13.2 


E 



Note. — The direct-acting steam pump (first item) has a duty of 
10,000,000 ft. lb. per 100 lb. of coal when used upon a boiler with 80 lb. 
per sq. in. gage pressure. This corresponds to a over-all efficiency of 
about 1.3 per cent. Figs. 194, 195, 196, 197 and 198 show how a set of 
values such as those above may be obtained. One pound of steam de- 



«*i 



176 



STEAM POWER PLANT AUXILIARIES 



[Div. 6 



livered to the engine is taken as the unit. The heat in both feed-water 
and steam above the feed-water temperature is considered. 



■1165 B.tM {1160 B.t.u. For Engine, 5 B.t.u. For Pump) 
''••---- . "p^ .Parti 

-Water 60° 



W\ x v 




\\<l\\\\\\\\\\Vv\\\V\\\;-. 
'Heoifer 

Heat Of Steam Above 60° Useol By Engine Anol Pump I 165 B.t.u. 
" " " " " Returned To Boiler 140 B.t.u . 

Net Heat Used IO?5B.t.u. 

E'Heotf Consumption Rehtive ToA=jf§%-= 0.868 

Fig. 198. — Power Feed Pump And Exhaust Heater. (It is here assumed that 
1,160 B.t.u. must be supplied per pound of steam required to drive the regular 
engine load as in the four preceding figures; but, on account of the additional 
engine load due to its having to drive the pump, the engine will now require more 
steam in the proportion of 1,165 to 1,160). 

204. The Definitions Of The Pump Designations Which 
Have Been Adopted In This Division To Denote The Three 
Different Types of Pumps are these: (A) Mechanically- 
driven pump, (Fig. 208) or pump which is mechanically 
driven from the engine. The drive may be either direct by 



Spring Mer Pivot, 
{Driving Motor 
\ Idler-, 



Water Relief Valve 



Pass 
6 Inlet 




Fig. 199.— Section Of Motor-Driven Tri- 



t ig i-ouer x aalrunip. 



a connecting rod, gears or belt; or indirect through a line shaft 
or other common forms of mechanical transmission. {B) 
Motor-driven pump, or pump operated by an electric motor 
installed or used specially for the purp^'ttnd belted (Fig. 
199), chain driven, geared (Figs. 200 ancBtol) or direct con- 
nected (Fig. 202) to the pump. (C) Steam driven pump or 






Sec. 204] 



BOILER-FEEDING APPARATUS 



111 



Driving Motor K 

Pinion 0/7-, 

Motor Shaft 

Crank--. 

trunk-Shaft 



Feed Pipe To Boilers-^ 
Gate Valve- ^ X 
\ Suction Pipe ■ 




Spur Wheel On 'Spur Wheel On Intermediate Gear- 
Crank-Shaft ShaftfPinfon, On Opposite End Of 
Gear-Shaft, Drives Spur Wheel On 
Crank- Shaft) 

Fig. 200. — Driving Motor Geared To Boiler-Feed Pump. 



(Electric Motor Mounted On Frame Of Pump 
'"A 




Fig. 201. — View Of Crank-End Of A Motor-Driven Boiler-Feed Pump. 



Gate Valve— 



driving Motor 

\ Shaft Coupling 



-<—Feed-P/pe To 

Boilers 
Multi-Stage Centri- 
fugal Pump- 




Suction Pipe From Feed-Water Heater--' 

Fig. 202. — Centrifugal Boiler-Feed Pump Motor-Driven Through Direct Shaft 

Connection. 
12 



178 



STEAM POWER PLANT AUXILIARIES 



[Div. 6 



pump operated directly by steam from the boiler and inde- 
pendently of the engine. In small plants they will be recipro- 
cating pumps of either the simplex or duplex type (Fig. 203), 

/Globe Throttle Valve (For Operation Of Pump With- 

I Out Governor Control) 

• /Globe ValvesfBoth Open 

• ; When Pump Is Running .-Weighted Lever 

j \Under Governor Control) / Fezof Lim To ^^ 



-Governor Valve 



Pipe For Conveying Feed- 
Line Pressure To Governor 
Diaphragm-^ 

.-^Diaphragm 

Globe Valve- 




Fig. 203. — Direct-Acting Boiler-Feed Pump Equipped With Fulton Governor. 



In plants of over 500 h.p., they may be either reciprocating 
or centrifugal pumps (Fig. 204). 



Multistage Cen- ,-Fzed Pipe To Boilers 
trifugaT Boiler \ Steam Turbine-^ 



Pressure 
Governor--^ 




Fia. 204. 



^Suction Pipe From Feed- Water Heater -Fxhaust Pipe 

-Centrifugal Boiler-Feed Pump Driven By Steam Turbine Through Direct 
Shaft Connection. 



205. The Possible Losses Due To Inefficient Boiler-Feed 
Pumps Are, In The Average Plant, A Very Small Proportion Of 
The Total Losses. — The total coal required, directly or indi- 
rectly, for the boiler-feed pumps in a reasonably-well designed 



Sec. 206] BOILER-FEEDING APPARATUS 179 

and operated plant of medium capacity is not liable to exceed 
more than 1 or 2 per cent, of the total coal burned. In a very 
small, inefficient plant, the proportion of the coal required for the 
boiler-feed pumps may in exceptional cases be as great as 10 
per cent. 

206. The Most Economical Type Of Boiler-Feed Pump And 
Drive Therefor Depend On Local Conditions. — Whether the 
plant is condensing or non-condensing, its horsepower capacity, 
the character of its load and the fluctuations thereof, the oppor- 
tunity to utilize exhaust steam, and other special and financial 
conditions may be factors. 

207. The Actual Cost Of Operation Of Any Boiler-Feed 
Pump Cannot Be Based Merely On The Efficiency Of The 
Pump Itself. — Nor can a comparison of the actual operation 
costs of boiler-feed pumps of the different types be based 
merely on the performance of the pump. The economic rela- 
tion of the other components of the plant wherein the pump is 
to be installed must be considered. Whether or not the ex- 
haust steam from a steam-driven feed pump can be utilized 
for boiler-feed-water heating or for building heating may be a 
determining factor. 

208. At Least One Reciprocating Steam-Driven Boiler- 
Feed Pump Should Be Installed In Every Plant. — If there is 
only one boiler-feed pump in a plant, it should be steam driven 
and preferably direct-acting. The reason for this is, as is 
explained in detail elsewhere in this Div., the inherent reli- 
ability of the direct-acting steam-driven pump, due to its sim- 
plicity and the fact that there are no links, except a steam line, 
between it and the boiler. Another advantage is that a steam- 
driven feed pump can, if there is steam in the boiler, be operated 
whether or not the engine is running. 

209. A Motor -Driven or a Power-Driven Boiler -Feed 
Pump Is Always Most Economical In A Non-Condensing 
Plant (if no live steam, in addition to exhaust steam, is required 
for building heating). The reason is that a non-condensing 
engine of itself will always (Sec. 240) furnish much more than 
sufficient exhaust steam to heat the feed water. Both the 
power-driven and the motor-driven pumps require considerably 
less coal for their operation than does a steam-driven pump. 



180 



STEAM POWER PLANT AUXILIARIES 



[Div. 6 



Hence, under these conditions, all the heat in the exhaust steam 
from a steam-driven pump would be wasted. (It requires, 
under ordinary non-condensing conditions, approximately 
14 per cent, or J<f of the exhaust steam from the engine to heat 
the feed water up to 212 deg., which is ordinarily the highest 
feasible feed-water temperature. The remaining % or 86 
per cent, of the exhaust steam from the engine is wasted into 
the atmosphere). 

210. In Any Plant Where Low-Pressure Live Steam In 
Addition To Exhaust Steam Is Required For Building Or Other 
Heating, A Direct-Acting Steam-Driven Pump Will Ordi- 



Exhaust To Atmosphere-., 



Back Pressure Yalvz 



V | I B 

Hive Steam From p% 
Boilers LJ 



y///////////w///Ff 




Steam Separator 




p»==u!\ 



Exhaust To-' 

Atmosphere 

"*"— Condenser 

-Motor-Driven Air 
Pump 

'Electric Motor 

-Motor-Driven Conden- 
sate Pump 



Steam Driven Circulating Pump-' 



Fig. 205. — Condensing Plant Equipped With Motor-Driven Boiler-Feed 
(Cochrane Heater.) 



Pump. 



narily Be Preferable because of its simplicity, reliability, and 
low first cost. Its steam consumption under these conditions 
is of minor importance because all of the heat in its exhaust 
steam is utilized for building or other heating. For building 
heating, exhaust steam is nearly as effective as live steam. 

211. The Use Of Some Non-Condensing Steam -Driven 
Auxiliaries Is Ordinarily Economical In Condensing Steam 
Power Plants (Fig. 205). — In condensing plants there is, as a 
rule, no exhaust steam available from the main engines for 
feed-water heating. When no economizer is used, the auxil- 
iary drives should be so proportioned that there will be just 



Sec. 212] 



BOILER-FEEDING APPARA TUS 



181 



enough exhaust steam from them to heat the feed-water to 
210 deg. fahr. This condition gives a maximum of economy 
as practically all of the energy of the steam delivered to the 
auxiliaries is then effective either as mechanical energy or heat. 
When an economizer is used in addition to an exhaust heater, 
some other feed-water heater discharge temperature such as 
150 deg. fahr. may prove economical. Both motor-driven and 
steam-driven pumps are often installed under such conditions. 
Then the operator uses whichever pump or pumps the exhaust 
steam from which will give the feed-water temperature from 
the exhaust heater which has previously been found to be most 
economical. 

212. An Automatic Exhaust Steam Heat Balance System (Fig. 
206) has been devised for maintaining a perfect balance between 



-Eleetric Control Wins 




Exhaust 
Condenszr Unit- 



Fig. 



206. — Diagram Of Plant With Condenser, Feed-Water Heater And Automatic 
Arrangement For Maintaining Heat Balance. 



the exhaust steam available and that needed for feed-water 
heating. When this system is used, most, or all, of the auxil- 
iaries are motor-driven and an auxiliary non-condensing- 
engine-generator unit, B, is provided to supply the auxiliary 
motors C and F with electrical energy. Then this auxiliary 
generator is interconnected with the main generator through a 
motor-generator set G, as shown in Fig. 206. Under normal, 



182 STEAM POWER PLANT AUXILIARIES [Div. 6 

full-load conditions, the motors driving auxiliaries take all of 
their electrical energy from the auxiliary generator, in which 
case it will then produce just enough exhaust steam to heat the 
feed water up to 212 deg. But, if due to change of load, the 
amount of exhaust steam supplied by this auxiliary engine- 
generator unit becomes more than sufficient to heat the feed 
water, then a portion of the electrical energy which the auxil- 
iary motor drives take is shifted over to the main generator E. 
This shifting of the electrical load is effected through the motor- 
generator G. 

Explanation. — This motor-generator is actuated by an electrical- 
contactor pressure valve H on the feed-water heater A. When there is 
a surplus of exhaust steam in the feed-water heater, the valve contactor 
operates in such a way that the motor-driven auxiliaries take a greater 
proportion of their energy from the main generator. When there is 
insufficient exhaust steam in the heater, the valve contactor operates in 
the other direction, causing the auxiliary motors to take more of their 
energy from the auxiliary generator. This throws a greater load on the 
auxiliary generator and results in the production of more exhaust steam 
by the auxiliary engine-generator unit. By means of this automatic 
arrangement, a practically-perfect heat balance always obtains. This 
results in the highest possible economy of operation. 

213. The Most Economical Boiler -Feeding Equipment For 
A Non-Condensing Plant Which Includes An Extensive 
Heating System For Winter Service is (Fig. 207) a steam- 
driven pump for operation during the heating season and a 
motor-driven or power-driven pump for operation when the 
heating system is out of service. 

Explanation. — If the heating system requires more steam in the 
winter than the main engine furnishes as exhaust, some live steam must 
be drawn through a reducing valve for heating purposes. A steam 
driven pump acts as a reducing valve and then furnishes power as a 
sort of by-product. The cost of the power, in this case, for pumping 
will be negligible. On the other hand, in the summer when the heating 
system is not in use there will be a great surplus of exhaust steam from 
the main engines alone. When this is true the motor driven or mechanic- 
ally driven feed-pump has the economic advantages as shown in Sec. 217. 

214. Mechanical-Drive For A Boiler -Feed Pump Is, Con- 
sidering The Feed-Pump Independently, Generally More 
Efficient Than Electric-Drive (Figs. 199, 200, 202, and 208). 



Sec. 214] 



BOILER-FEEDING APPARATUS 



183 



-Back Pressure. Votive 




^ j ^^^fF^m 



fr t$%. ^ |^E^ ^y-^^ 



Fig. 207. — Non-Condensing Plant Equipped With Steam-Driven and Motor-Driven 
Boiler Feed Pumps. 



.■Countershaft 

'driving Pulley 
Driving Belts- 




Crank- 



mmmz0Mz77zm7, 



Fig. 208. — A Mechanically-Driven Boiler Feed-Pump. (For boilers carrying steam 
pressure under 85 pounds per sq. in.) 



184 STEAM POWER PLANT AUXILIARIES [Div. 6 

This is due to the fact that, under average conditions, the total 
mechanical losses between the main engine and a mechanically- 
driven pump are less than the total mechanical and elec- 
trical losses between the main engine and a motor-driven 
pump. It should be understood, however, that this rule 
holds only under favorable conditions. If, in order to drive a 
feed pump mechanically, it is necessary to transmit the power 
through long line shafts, through many belts, and through 
right-angle-turn belt transmissions, then the higher theoretical 
efficiency of the mechanical drive will vanish. Furthermore, 
the efficiencies upon which the following example is based, 
relate to mechanical transmissions which are well installed and 
maintained. Line shafts out of alignment, worn bearings, and 
similar conditions may result in material decrease in efficiency. 
It should be noted, then, that while the over-all efficiency of 
the electrical drive will remain practically constant throughout 
the life of the drive, because the elements which affect it are 
the generator, electrical transmission and motor efficiencies, all 
of which are unaffected as time progresses, the efficiency of the 
mechanical drive may decrease, due to use. 

Example. — Suppose that power is transmitted from the main engine 
to a mechanically operated feed-pump through two successive belt con- 
nections — the first being from the engine-shaft to a line-shaft, and the 
second from the line-shaft to the pinion-shaft of the pump. Also, sup- 
pose the transmission efficiency of the belting is 97 per cent., of the fine 
shaft 96 per cent., and of the pump gearing 96 per cent. The overall 
efficiency of the mechanical drive is, then, 0.97 X 0.96 X 0.96 = 89.4 
per cent. Suppose, also, that power is transmitted from the main engine 
to a motor-operated feed-pump which is working under the same service 
conditions as the mechanically-driven pump. Then, assuming a gener- 
ator efficiency of 93 per cent., a transmission line or wiring efficiency of 
95 per cent., a motor efficiency of 85 per cent., and a pump-gear efficiency 
of 96 per cent., the overall efficiency of the motor-drive is 0.93 X 0.95 X 
0.85 X 0.96 = 72 per cent. Thf conomical advantage of the mechan- 
ical over the motor drive is, thei. , 3, represented by an efficiency differ- 
ence of 89.4 - 72 = 17.4 per cen 

\ il 

215. Motors For Driving Fee I Pumps should be of enclosed 
or semi-enclosed types if the pumps are installed in a dusty 
boiler room or in any other dusty place. The motor should 
preferably be of the adjustable-speed type so that the water 



Sec. 216] BOILER-FEEDING APPARATUS 185 

may be pumped into the boiler at the same rate as that at 
which it is evaporated. The rate of evaporation varies with 
the load. 

Note. — An Adjustable Speed Motor is one the speed of which can 
be varied over a considerable range and when once adjusted remains 
practically unaffected by the load. Examples are shunt wound, lightly- 
compound-wound d.c. motors. A varying-speed motor is one the speed 
of which varies with the load, such as a d.c. series or heavily-compound- 
wound motor or an a. -c. wound-rotor slip-ring induction motor. Since 
adjustable-speed motors, capable of sufficient speed variation for efficient 
boiler feed service, are not ordinarily obtainable in the smaller capacities, 
it is necessary to use varying-speed motors for these small-capacity 
applications. 

216. If A Constant-Speed Motor Is Used On A Boiler 
Feed Pump either the water feed must be intermittent, which 
is undesirable, or, if the motor continues to operate at constant 
speed, a part of the feed water must be by-passed through a 
by-pass valve. Where a by-pass valve is used, the motor may 
operate continuously at constant speed and little or much of 
the water it pumps be admitted to the boiler by controlling 
the by-pass valve as occasion requires. This by-passing is 
very uneconomical because then all the water handled must 
be pumped against boiler pressure. Then the energy imparted 
to the portion of the water which is not fed into the boiler is 
wasted. This situation may be practially corrected in the 
larger plants by installing two feed pumps, each of one-half 
the capacity necessary for total requirements. 

Note. — Gear Drive Is Preferable To Belt Drive, because of 
high-cost maintenance. Feed pumps are frequently installed in out-of- 
the-way corners where it is difficult to make prompt repairs on belts. 
The belt may slip off or break when suc T i an accident can be least afforded. 

217. The Economy Of A ichanically-Driven Boiler- 
Feed Pump Which Operates I A Constant Speed is affected 
adversely where the load oi the plant fluctuates widely. 
This is due to the fact that the water horsepower output of 
the pump cannot be varied economically in response to the 
fluctuations of the load. 



186 



STEAM POWER PLANT AUXILIARIES 



[Div. 6 



Note. — A mechanically-driven pump for boiler-feeding, and its driver, 
should (Sec. 231) be designed to meet the maximum requirements of the 
boilers. The quantity of water delivered to the boilers may be regu- 
lated (Fig. 209) in accordance with load variations, by means of a by-pass. 
But with this method of control, a large proportion of the power con- 
sumed by the pump is wasted. Although the quantity of water passing 
into the boilers through the check-valves in the delivery pipes may be 
diminished, the pressure at the by-pass connection to the discharge pipe 
will only be slightly less than the boiler pressure. The pump will, there- 
fore, still be discharging at its full capacity against almost full boiler 
pressure. The average economy of a mechanically-driven boiler-feed 
pump (considered as an independent unit), operating under the most 
extreme conditions of load-fluctuation which might prevail, would 
nevertheless be generally superior to that of a steam-driven pump oper- 
ating under like conditions. 



Belt-Connection To Line-Shaft- 
Countershaft---^^ 




Outs/ae-Center- 

Po/cked Plunger 

Pump.. 



Fig. 209. — Mechanically-Driven Boiler-Feed Pump Furnished With By-Pass. 



Example. — Suppose a steam-driven pump, which consumes 175 lb. 
of steam per h.p. per hr., develops 4 h.p. while feeding a fully loaded set 
of boilers. Then the quantity of steam required to operate this pump 
will be 175 X 4 = 700 lb. per hr. Also, suppose a mechanically-driven 
pump, operated by main-engine power which is transmitted to the pump 
on a steam consumption of, say, 30 lb. per hr., develops 4 h.p. while 
feeding a similar set of boilers. Then the quantity of steam required 
to operate this pump will be 30 X 4 = 120 lb. per hr. If the load on 
each set of boilers is diminished one-half, then the power required to feed 
tl$em will be 4 -¥■ 2 = 2 h.p. The steam pump need then be run at only 
one-half its former speed. Hence, it will develop no more than the 
requisite 2 h.p. during the period of half load. Its steam-consumption 
will, therefore, be reduced to 700 -f- 2 = 350 lb. per hr. The mechanic- 



Sec. 218] BOILER-FEEDING APPARATUS 187 

ally-driven pump will, however, maintain its original speed. It will, 
therefore, continue to discharge at the rate of its full-load capacity. 
But only one-half of the water discharged will enter the boiler. The 
remaining half will be by-passed. All of .the water will, nevertheless, be 
discharged against the boiler pressure. Hence, the pump will continue 
to develop 4 h.p. during the half-load interval. Consequently, it will 
maintain its original steam rate of 120 lb. per hr. Its economical 
advantage over the steam pump will, therefore, be reduced from 
[(700 - 120) -f- 700] X 100 = 83 per cent, under full-load conditions to 
[(350 -120) -7- 350] X 100 = 66 per cent, under half-load conditions. 

218. The Saving Which May Be Effected By Substituting 
An Electrically-Driven For A Steam-Driven Feed Pump 

may be estimated as follows: The example is based on a 
non-condensing steam plant, which has a feed-water heater, 
and which operates twenty-four hours a day. It is assumed 
that the conditions are such, as is usually the case in a non- 
condensing plant, that the exhaust steam from the steam- 
driven pump, direct-acting or turbo-centrifugal, cannot be 
utilized effectively for feed-water heating or otherwise. 
(The engine alone in a non-condensing plant furnishes about 
six or seven times as much exhaust steam as can possibly be 
reclaimed for feed-water heating. Hence, in such a plant the 
exhaust steam from the feed pump represents pure waste.) 

Example. — Average load on plant, 50 kw. Assumed water rate for 
this simple engine plant, 50 lb. of steam per kw.-hr. Therefore, the steam 
consumption per hr. for this 50-kw. plant = 50 X 50 = 2,500 lb. of 
steam per hr. Hence, Gal. of feed water required per hr. = 2,500 -s- 8.3 = 
300 gal. per hr. (approx. ) . Pressure against which feed water is forced = 
125 lb. per sq. in. Head = lb. per sq. in. X2.31. Hence, for this plant, 
head = 125 X 2.31 = 290 ft. approximately. Work required per hour 
to force water into boiler = weight of water per hr. X head = 2,500 X 
290 = 725,000 ft.-lb. per hr. 

Now the average expected duty of an ordinary 4-in. -stroke steam- 
driven boiler-feed pump may be stated conservatively as 11,700,000 
ft.-lb. per 1,000 lb. of steam. Or, in other words, it may be assumed 
safely that for a steam boiler-feed pump in this 50-kw. plant: 

No. of ft.-lb. per lb. of steam = — Yqqq — = 11,700 /*.-Z6. 

As is evident from preceding statements, the pounds of steam con- 
sumed per hour in this plant to develop the 725,000-ft.-lb. required to 



188 STEAM POWER PLANT AUXILIARIES [Div. 6 

pump the 2,500 lb. of water against the 125 lb. per sq. in. steam pressure 

is: 

T , o . . j . 7 ft .-lb. to supply water per hr. 

Lb. of steam to drive steam pump per hr. = J - — - — ^— ^ — ^ 

ft. -lb. per lb. of steam 

= .. 1 '„ = 62 lb. of steam per hr. 

That is, a steam-driven pump for this plant would consume 62 lb. of 
steam per hr. It would consume 62 lb. of steam in pumping 300 gal. 
boiler-feed water. 

Now the steam required to operate an equivalent electrically-driven 
pump will be determined: A test, reported by the Midvale Machine 
Company, indicates that 360 gal. of boiler-feed water were pumped, in 
a plant similar to that under consideration, with an energy expenditure 
of 0.68 kw.-hr. by a Johns electrically-driven pump. The following 
determination will be based on the data of this test. 

In the 50-kw. plant under consideration, it is, as previously stated, 
assumed that there will be required 50 lb. of steam to develop 1 kw.-hr. 
Hence, to develop 0.68 kw.-hr. there would be required: 0.68 X 50 = 34 
lb. of steam. Now if 360 gal. of water were pumped in the test by 34 lb. 
of steam, the 300 gal. of water, required per hour in this plant would be 
pumped by 300 ^- 360 X 34 = 28.5 (approx.) lb. of steam. 

Now the saving in steam due to the use of the Johns electrically-operated 
pump will be the difference between the steam-pump steam consumption 
and the equivalent electrically-driven pump steam consumption. Thus: 

Steam pump requires per hr. (to pump 300 gal. feed 

water) 62 . lb. of steam 

Elec. driv. pump requires per hr. (to pump 300 gal. 

feed water) 28 . 5 lb. of steam 

Saving in steam per hr. due to use of elec. driven pump 33 . 5 lb. of steam 

Now the saving in dollars due to the use of the motor-driven pump can 

be determined: The boiler evaporation in a small plant, of the character 

of that under consideration, will be about 7 lb. of water for each pound of 

coal. Then, the coal required to evaporate 33.5 lb. of water (the amount 

which is saved, each hour, by the use of the electrically-driven pump) 

will be: 

lb. of water evaporated 33.5 . „ „ » , , 

f — - £— = — =— = 4.8 lb. of coal per hr. 

rate of evaporation 7 

For one month the coal saved, based on this 50-kw. load, would be 
30 days X 24 hr. X 4.8 lb. per hr. = 3,460 lb. per month. With coal 
costing $3.00 per ton in the bin, the saving per month would be: (3,460 X 
3.00) -v- 2,000 = $5.19. Or the saving per year would be, approxi- 
mately: 12 X $5.19 = $62.25. 



Sec. 219] 



BOILER-FEEDING APPARATUS 



189 



219. Table Showing General Advantages And Disadvan- 
tages For Boiler Feeding Of Power, Electric, and Steam 
Pumps When Considered Independently. 



A 

Mechanically driven 




Steam driven 



Advantages 



1. Simplicity. 



2. Low cost of equipment. 



3. Lowest theoretical cost 
of operation where non- 
condensing engine is used 
and pump is' belted direct. 

4. Cannot run away in 
case of accident. 



1. May be easy of con- 
trol. (Automatic control 
possible.) 

2. Fairly simple. 



3. More efficient than 
steam — low cost per year. 



4. Location where desir- 
able. 

5. Can be operated in- 
dependently of boiler and 
engine unit under some 
circumstances. 

6. In case of accident, 
has speed limit, i.e., will 
not "run away." 

7. Accurate heat balance 
possible with proper equip- 
ment. (Sec. 212.) 



1. Ease of control. 



2. Supply of exhaust 
for feed-water heat- 
ing in condensing plant. 

3. Low cost of equipment. 



4. Maximum reliability, 



Disadvantages 



1 . Poor regulation of 
water supply — extra water 
may be pumped and re- 
turned if by-pass is used. 

2. May be unhandy in 
location. 



3. Works only when main 
engine is running. 



1. Cannot be used with- 
out generator. (Unless 
Public-Service-Company 
power is available. *) 

2. In condensing plant 
increase loss of exhaust 
steam heat in condensing 
water unless pump is driven 
from auxiliary non-con- 
densing engine unit exhaust 
from which is used for feed- 
water heating. 



1. Great loss due to in- 
efficiency unless exhaust 
steam can be utilized. 

2. Possible self destruc- 
tion in case of feed-line 
breaking. 



* Where Public-Service-Company power is available this feature may be important. 



190 



STEAM POWER PLANT AUXILIARIES 



[Div. 6 






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Sec. 220] 



BOILER-FEEDING APPARATUS 



191 





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192 STEAM POWER PLANT AUXILIARIES [Div. 6 

221. A Turbine- Or Motor -Driven Centrifugal Pump 

(see Div. 4) affords, ordinarily, the best unit for regular 
operation for pumping boiler feed water for plants of capaci- 
ties exceeding about 500 h.p. A 500-h.p. plant is equivalent 
to a feed-water requirement of about 50 gal. per min. or 3,000 
gal. per hr. (However, in every case there should be a steam 
direct-acting stand-by pump.) The centrifugal pump is the 
best for this service because it will, in the long run, prove the 
most economical. It has the advantage that the discharge 
from the pump to the boiler may be throttled down or opened 
as desired without the considerable loss of energy which 
results from by-passing. 

Note. — The pressure developed by a centrifugal pump which is oper- 
ated at normal speed can never exceed a certain maximum. Further- 
more, if the feed line from the pump to the boiler should break, thus 
reducing the head against which the pump is forcing water to practically 
zero, the centrifugal pump will not "run away," but it will continue to 
operate at practically constant speed. Its power consumption will be very 
low when it is pumping against zero head. Again, if the valve in the 
discharge pipe in a centrifugal pump is closed the pump may continue 
to turn at its normal speed (Sec. 171) without developing an excessive 
water pressure. In such a case the water is merely churned around 
within the casing. 

222. The Efficiency Of The Centrifugal Pump remains, 
with slight repair, nearly constant throughout its life because 
there is practically nothing about it except two simple bearings 
to wear out. Obviously, where gritty water is being pumped 
through there will also be wear on the impellers or blades, but 
gritty water is not used for boiler feed. On the other hand, 
the efficiency of any plunger type or piston pump may decrease 
decidedly as the pump becomes older, due to leaky valves, 
pistons, and worn rods. This is true particularly of the single 
or duplex steam pump. The water rate of such a steam pump 
after a year or so of service and insufficient maintenance may 
be twice its initial water rate. 

223. The Centrifugal Pump Has No Valves Which Re- 
quire Re -Grinding. — Unfortunately, the valves of any plunger 
or piston pump do not usually receive the attention which 
they should have. With the steam pump, if the valves 
become leaky the operator may merely "give her more 



Sec. 224] BOILER-FEEDING APPARATUS 193 

steam." Thus, the required water may be pumped, but 
uneconomically. Such losses are difficult to locate because 
the steam requirements of the boiler feed pumps are such a 
small proportion of the total steam requirements of the plant. 

Note. — The Water Rate Of A Turbine For Driving A Small 
Centrifugal Pump will be from 38 to 43 lb. of steam per brake h. p. hr. 
This consumption does not increase materially as the age of the turbine 
increases. 

Note. — The Mechanical Efficiency Of A Centrifugal Pump 
(the capacities range from 50 gal. per min. and up) will be from 50 to 
60 per cent. For the larger centrifugal pumps operating under favorable 
conditions efficiencies as high as 81 per cent, have been obtained. 

224. One Disadvantage Of Centrifugal Pumps for boiler- 
feeding is the fact that if the feed water is very near its boiling 
point, the action of the pump may vaporize it entirely within 
the pump casing (see Sec. 157). If this happens, the pump 
will not work as it depends on the action of the runner on a 
liquid. On the other hand, a plunger pump will handle water 
at any temperature as long as there is pressure enough to 
deliver the water to the pump cylinders. Moreover a cen- 
trifugal pump cannot be run at all if in poor condition, on 
account of its high speed. If there is any damage to shaft 
or runner, the pump must usually be shut-down and com- 
pletely overhauled. It is for these reasons that a centrifugal 
pump is not recommended in this Div. for a stand-by pump. 

225. Power Boiler -Feed-Pump Sizes For Various Boiler 
Horse Powers as taken from The Goulds Mfg. Go's, catalogue 
are given in the two tables which follow. The tabulated 
values indicate the water supply required by the boiler based 
on the A. S. M. E. standard rating (Sec. 229) of 34>^ lb. of 
feed water per boiler horse power hour from and at 212 deg. 
fahr. A surplus of 25 to 50 per cent, pump capacity is 
recommended. See Sec. 228 for methods of computing boiler 
feed-water requirements. 

226. Table Showing Boiler-Feed Capacities Of Single- 
Acting Triplex Power Pumps. Goulds Mfg. Co. (See preced- 
ing Sec). The capacity of a double-acting simplex pump is 
approximately 0.66 times of that tabulated for the same speed 
and cylinder dimensions. The capacity of a double-acting 

13 



194 



STEAM POWER PLANT AUXILIARIES 



[Div. 6 



duplex pump is 1.33 times that tabulated for the same speed 
and cylinder dimensions. 



Rated capacity of 

boilers, 

horse power 


Feed water at 
212°F., 


Size of pumps, 


Revolutions 


Gallons per 


inches 


per minute 




minute 






30 


2.15 


m X 2Y 2 


30 


50 


3.59 


2X3 


31 


1Q0 


7.17 


2V 2 X 4 


30 


** T5V 


10.75 


3 X 4 


31 


200 


14.34 


3^X4 


30 


400 


28.7 


4 X 6 


31 


800 


57.4 


5 X 8 


30 


1200 


86. 


6 X 8 


31 


1600 


115. 


7 X 8 


30 


2000 


143.4 


7 X8 


30 


2750 


196. 


8 X 10 


31 


4000 


286. 


9 X 12 


30 


5000 


358. 


10 X 12 


30 



227. Table Showing Boiler-Feed Capacities Of Multi- 
stage Centrifugal Pumps. Goulds Mfg. Co. (See Sec. 225). 





Feed 


Size of 






Feed 


Size of 




















power of 
boilers, 
rated 


water at 


pipe 


Revo- 




water at 


pipe 


Revo- 


212°F., 
gallons 


dis- 
charge, 


lutions 
per 


power of 
boilers, 
rated 


212°F., 
gallons 


dis- 
charge, 


lutions 
per 


capacity 


per 

minute 


pipe 
inches 


minute 


capacity 


per 

minute 


pipe 
inches 


minute 


700 


50.20 


2 


3500 


2800 


200.5 


4 


2500 


850 


60.92 


2 


3500 


3150 


225.9 


4 


2500 


1000 


71.60 


2 


3500 


3500 


250.8 


4 


2500 


1200 


86.00 


2 


3500 


3850 


275.8 


4 


2500 


1500 


107. 50 


2 


3500 


4200 


301.0 


4 


2500 


1750 


125.30 


2 


3500 


4900 


351.0 


5 


2200 


2000 


143.33 


3 


3100 


5600 


402.0 


5 


2200 


2100 


150. 50 


3 


3100 


6300 


452.0 


5 


2200 


2450 


175.50 


3 


3100 


7000 


502.0 


5 


2200 



228. There Are Two Methods Of Estimating Feed-Water 
Requirements Of A Power Plant. — One is based on the rating 
of the boilers in the plant. The other is based on the actual 
steam consumption of the engines and auxiliaries or devices 



Sec. 229] BOILER-FEEDING APPARATUS 195 

for which the steam is generated. Which method should be 
used in any case will be determined by conditions. Probably 
the second method, that based on the actual steam consump- 
tions, is the more accurate. But, in a plant in which steam 
is used only for power generation, if the boiler capacities are 
proportioned rationally in relation to the units which they 
supply, both methods should give approximately the same 
results. In ascertaining the feed-water requirements for a 
power plant it may be wise to make an estimate by each of the 
methods, compare the results as a check, and then take for a 
working basis the one which is the larger. Where a boiler 
plant generates steam for heating only, that is, where there 
are no steam-consuming units, such as pumps and engines, it 
is obvious that then only the first method, that based on the 
boiler rating, is applicable. 

229. In Determining Feed-Water Requirements On The 
Basis Of The Boiler Rating the accepted water-rate equiva- 
lent of a boiler horsepower (boiler h.p.) is utilized. The 
equivalent is this: It was recommended by the American 
Society of Mechanical Engineers in 1899 that the evaporation 
of 34.5 lb. of water per hr. at 212 deg. be taken as the equiva- 
lent of 1 boiler h.p. This equivalent is now universally 
accepted as standard in the United States. Hence, the process 
of determining the water required to feed a boiler is this: 
(1) Ascertain the total h.p. rating of the boiler or boilers in 
question. (2) Multiply this total h.p. rating by 34.5 which 
will give the number of pounds of water required per hour 
when the boiler is operated at rated capacity. (3) Now due 
to the fact that the pump must occasionally raise the water 
level and pump more than 34.5 lb. per hr., the value obtained 
in this manner should be increased by 30 to 50 per cent. 

In fact, a boiler feed pump is usually selected on the basis 
that it will deliver 45 to 50 lb. of water per hr. for each rated 
boiler h.p. The operations above applied may be expressed 
in a formula, thus : 

(71) Lb. of water per hr. = W wh X Pb* p 

Wherein : W wh = the lb. of water per boiler h.p. hr. upon which 
the estimate is based. This value may vary from 45 to 50. 



196 STEAM POWER PLANT AUXILIARIES [Div. 6 

The value of 45 lb. per hr. is conservative and may ordinarily 
be assumed. V B h P = the total rated boiler h.p. of the boiler 
or boilers which are to be fed. Now since there are 8.34 lb. 
of water in a gallon, it follows that : 

(72) Gal. required per hr. = W »» * P *»p 

o.o4 

Now if W W h be taken as 45, then 

(73) Gal. required per hr. = 45 X P B h P -*- 8.34 = 5.4 P Bhp . 

which is the accepted working formula. Where Wu* is taken 
as 50 lb.: 

(74) Gal. required per hr. = 6 X P*Ap. 

Example. — A boiler has 500 rated h.p. What should be the capacity 
of the feed-water pump to supply it? Solution. — Base the estimate on 
45 lb. of water per rated h.p. hr. Then substitute in the above formula, 
thus: Gal. required per hr. =5.4 XPiihp = 5.4 X 500 = 2,700. 

Hence, a pump capable of delivering at least 2,700 gal. of water per 
hour should be installed. 

230. In Determining The Feed-Water Requirements Of A 
Power Plant On The Basis Of Its Steam Consumption, the 

process is this: (1) Ascertain, either from manufacturers' 
guarantees, or by using a table of water rates, the pounds per 
hour of steam required for the engine or principal units. (2) 
Similarly determine the pounds of steam required per hour by 
the auxiliaries. Then disregarding radiation, leakage, steam 
required by the whistle, and other losses : 

(75) Total weight of water required per hr. = (1) + (2) 

To allow for the radiation, leakage, whistle loss, and to 
provide some capacity for forcing and for recovering the 
water level in case it is lost, the value obtained by the equation 
just above should be increased by 25 per cent. 

Example. — What will be the probable feed-water requirement of a 
plant which operates a 50-h.p. high-speed condensing engine and 10 h.p. 
of non-condensing auxiliaries? Solution. — First determine the water 
consumption of engine and auxiliaries. From a table of water rates it is 
found that a 50-h.p. high-speed condensing engine will have a water rate 
of about 22 lb. per h.p. hr. Hence, its total steam consumption will be: 
(50 X 22) = 1,100 lb. per hr. The water rates of the small auxiliaries 



Sec. 231] BOILER-FEEDING APPARATUS 197 

will probably be 200 lb. per h.p. hr. Hence, the total auxiliary steam 
consumption will be: (10 X 200) = 2,000 lb. of steam per hr. Steam con- 
sumption of engine and auxiliaries, then, is: (1,100 + 2,000) = 3,100. 
Multiplying this by 1.25 to allow for losses and forcing, thus: (3,100 
X 1.25) = 3,875 lb . This is the total weight of water required per 
hour. To reduce this to gallons, divide by 8.3, thus (3,875 4-8.3) = 
467 gal. per hr. 

231. Increased Feed-Pump Capacity Is Necessary If The 
Modern Large -Plant Practice Of Forcing The Boilers Is 
Followed. — In large power plants where automatic stokers 
can be used, particularly if the plant is situated in a city, 
boilers are " forced" so that their output much exceeds the 
nominal evaporation of 34.5 lb. per rated boiler h.p. per hr. 
by possibly 150 to 200 per cent. A forced boiler is not as 
efficient as one which is being worked conservatively. This 
is because that, when a boiler is forced, a larger proportion of 
the heat of the coal is wasted in the flue gases than if the boiler 
is not forced. That is, the flue gas temperatures in the smoke 
stack will be higher in the case of the forced boiler. This is 
equivalent to a loss. But in spite of the fact that the boiler 
efficiency is decreased when the boiler is forced, it usually 
works out in the larger power plants that it is more economical 
to force the boilers than it would be to pay the additional 
fixed charges on boiler investment, maintenance and real 
estate that would be involved if sufficient boiler capacity were 
installed to insure operation on the basis of an evaporation of 
34.5 lb. of water per rated boiler h.p. per hr. 

Note. — The Life Of A Boiler Which Is Being Forced may not, 
unless the forcing is extremely excessive, be less than that of a boiler 
which is not forced. It is essential, however, that a forced boiler be 
provided with purified feed water, otherwise scaling and blistering 
difficulties are bound to occur. 

Note. — In Practice, Boilers In Large Plants Are Now Often 
Forced To 150 To 200 Per Cent, of the A. S. M. E. rating for normal 
operation and at peak load they may be forced to 300 per cent, rating. 
That is, for each rated boiler h.p. (on the A. S. M. E. basis of an evapora- 
tion of 34.5 lb. of water per hr.) a boiler which is being forced to 150 per 
cent, of its rating will then evaporate: 34.5 X 1.5 = 51.75 lb. of water per 
hour. Similarly, if a boiler is being forced to 200 per cent, rating it 
will evaporate 69 lb. of water per rated boiler h.p. per hour. At peak 
load periods when forced to 300 per cent, rating, the evaporation may be 



198 



STEAM POWER PLANT AUXILIARIES 



[Div. 6 



Pipe Connection 
To Feed Line- 



Diaphragm 



103.5 lb. of water per rated boiler h.p. per hour. So it is evident that 
the rule given in one of the opening paragraphs of this section for com- 
puting feed-pump capacity on the basis of rated boiler h.p. may have 
to be modified materially if the pump is to be used in the plant where 
the boilers are forced. In such plants the safer procedure is to determine 
the actual steam consumptions per hour of the prime movers and of all 
of the auxiliaries and base the feed-pump rating on this total steam con- 
sumption. Furthermore, in estimating boiler-feed-pump capacities, 
ample allowance should be made for future additions to the boiler 
equipment, if any are contemplated. 

232. Pump Governors On Direct-Acting Steam Pumps In 
Boiler -Feed Service (Fig. 210) operate in conjunction with 

feed-water regulators. (See the 
author's Steam Boilers.) The 
function of the pump governor is 
to maintain a constant pressure 
in the feed line. It does this by 
moderating the speed of the 
pump, or shutting it down alto- 
gether, when the feed-water reg- 
ulator diminishes the openings 
through the feed valves or closes 
them entirely. If no governor 
were used to regulate the speed 
of the pump in response to the 
feed-water regulator's adjust- 
ment of the feed-valves, the 
pump might build up a pressure 
in the feed-line powerful enough 
either to force an excess quan- 
tity of water into the boilers 
through the partially closed feed- 
valves or to burst the piping. A 
properly working pump governor 
controls the movement of the pump piston or pistons so as to 
constantly maintain a pressure in the feed line just enough 
greater than the boiler pressure to insure a positive flow of 
the water into the boilers against the steam pressure. 

Explanation. — The Fulton pump governor (Fig. 211) is connected 
into the live steam supply at B and to the steam end of a direct-acting 




Fig. 210. — Sectional Elevation Of 
Fisher Pump Governor For Boiler- 
Feed Pumps. 



Sec. 232] 



BOILER-FEEDING APPARATUS 



199 



steam pump at A. The small pipe E is connected to the feed line so that 
the under side of diaphragm D is subjected to feed water pressure. The 
upper side of D is subjected to live steam pressure (or approximately 
boiler-pressure) through the passage C. The vertical stem of valve V is 
acted upon by weight W at one end and 
diaphragm D at the other When the 
feed line pressure is less than boiler pres- 
sure, both the weight and the diaphragm 
tend to open the balanced valve V. Then 
steam flows freely through the governor 
from B to A and operates the pump at 
full speed. The feed line pressure is built 
up by the pump until the weight W is 
lifted and valve V closed by the pressure 
on the under side of the diaphragm. The 
weight may be adjusted to open the valve 
at any desired pressure. 

Example. — Suppose the boiler pressure 
is 100 lb. per sq. in. and a pressure of 
110 lb. per sq. in. in the feed line is satis- 
factory for delivering water against the 
pressure of the boiler. Therefore when 
the pump is working at the proper rate, 
there will be 110 lb. per sq. in. on the 
under side of the diaphragm (D Fig. 211) 
and 100 lb. per sq. in. on the upper side. 
The weight W is set so as to overcome the 
force of this difference in pressure of (110 — 100) or 10 lb. per sq. in. 
Therefore when the difference in pressure is a little less than 10 lb. per 
sq. in., the weight opens the valve V. When the difference in pressure 
is a little more than 10 lb. per sq. in., the diaphragm closes the valve. 
In this way a pressure difference just sufficient to feed the boiler is 
maintained. 




■Dhphracfr, 



■Pipe Leading 
From Feed- Line 



Fig. 211. — Sectional Elevation 
Of The Fulton Governor For 
Boiler-Feed Pumps. 



[n 



£$& &*vs 




Drip Line-' 



FeecfL/'ne- 



Ci'f/et- 
Fig. 212. — Section Of Horizontal Piston Type Pump Governor. 

Note. — In The Horizontal Type Of Pump Governor (Fig, 212), 
the piston P takes the place of the diaphragm and a spring S takes the 
place of the weight as described above. The piston, however, is acted on 
by the feed line pressure only and so communicates this full pressure to the 
spring. The tension on the spring is adjusted by means of two thumb- 
screws T. 



200 



STEAM POWER PLANT AUXILIARIES [Div. 6 



233. The Fisher Pump Governor (Figs. 210 and 213) is 
similar to the Fulton governor in operation. One advantage 

Pipe Conveying Steam Pressure 
To Uno/er Sio/e Of Diaphragm- 

] Pipe Conveying Feed-Line Pressure 7b-* 




'Suction Pice From Feed- Water Heater 
Fig. 213. — Direct-Acting Boiler-Feed Pump Equipped With Fisher Governor. 



Steam Mam--'^ 

From Main Steam Line To Under Side 
Of Diaphragm- 



/Main Feed Line To Boilers 

\ From, Main Feed Line To Upper 



Side Of Diaphragm 



Steam 
Supply 

To 
Pump-, 




Fig. 214. — Direct-Acting Boiler Feed Pump Equipped With Kieley Governor. 

of the Fulton design shown is that it uses only one stuffing 
box where the Fisher design requires two. One advantage 



Sec. 234] BOILER-FEEDING APPARATUS 201 

of the Fisher design is that the steam pressure can be shut off 
from the diaphragm chamber for inspection and repair. The 
Kieley governor (Fig. 214) uses a spring in connection with a 
diaphragm in chamber D. 

Note. — Pump governors for maintaining constant-pressure are some- 
times used on turbine-driven centrifugal boiler-feed pumps (Fig. 204). 
Their function is merely to save steam, as there is little danger from the 
over-pressure of a centrifugal pump. A centrifugal over-speed governor 
may be provided on the same unit as a constant-pressure governor. 

234. A Water-Relief Valve Must, Where A Feed-Water 
Regulator Is Used, Be Installed On A Constant-Speed Crank- 
Action Feed Pump. — The water-relief valve (Fig. 199) is 
merely a special-type safety valve. Crank-action feed pumps 
usually run at fairly-constant speed (Sec. 217) and are always 
pumping about the same amount of water. Hence, if the 
feed-water regulator partially or wholly closes the feed-water 
line to the boilers, stalling of the pump or damage to the pump 
and its accessories are liable to result unless a water-relief 
valve is provided to automatically by-pass the surplus water. 

Note. — A Water-Relief Valve Should Be Provided In The By- 
Pass On Every Reciprocating Power Feed Pump as a safety measure, 
whether or not a feed water regulator is employed. This is to prevent 
damage if the feed- water line to the boilers is closed accidentally. 

235. The Most Common Troubles Of Pump Governors 
And Their Causes And Remedies are as follows : 

1. Blows Steam Around Valve Stem. Should be entirely re- 
packed with fine packing and lubricated with cylinder oil and graphite. 
Screwing up the packing gland to stop steam leaks is likely to make too 
much friction before the gland is tight. 

2. Too Sluggish — gives too much variation in feed-line pressure. 
Friction in the movement usually gives this effect. Sometimes the 
spring used is too stiff for the pressure in governors of the spring type. 
See if the valve stem slides freely. If it does not, the friction must be 
located and then remedied by polishing and lubrication. Sometimes a 
stuffing box is too tight or the packing old and stiff. A weaker spring 
gives less variation of feed-line pressure. 

3. Gives Constant Pressure Too Low Or Too High. Adjust 
weight or spring thumb screws. Increase spring tension or weight 
leverage for more pressure. 

4. Does Not Shut Off — gives excessive feed-line pressures as shown 
by gage or by creeping or other signs of overpressure in pump. 



202 



STEAM POWER PLANT AUXILIARIES 



[Div. 6 



(a) Friction in stem or piston. Remedy as explained before. 

(6) Damaged diaphragm. Remove this member. A high-grade rub- 
ber packing re-inforced with several layers of fabric may be used for 
diaphragms but plain rubber is not suited for the purpose. An extra 
diaphragm should be ordered from the manufacturer and kept on hand. 

(c) Valve does not seat. Examine seat for scores and corrosion. If 
it seems to be in fair condition grind with grinding compound and see if 
a clean face can be obtained. If scored too deeply to be ground clean, 
the valve must be re-finished on a lathe. In re-finishing, the angle of 
the face and span between the two faces must be accurately retained. 
After finishing, the valve should be "ground in" and all grinding com- 
pound removed before re-assembling. 

236. Automatic Apparatus For The Feeding Of Boilers 
With Hot-Water Returns from heating systems are of two 



Counter-weight For Closing/ Steam. Y&jM;. 

~ m- ~™ u~u,„ >''' Inlet For Returns 
1 r Steo/m VwvejfYom Heating Sys- 

Cocks Hmi ; ,J&&.-Lubr/cator 



Vent 
To Poof-. 




Suction Connection 
To Stand Pipe...... 



' Outlet To, 

Stand Pipe 



Fig. 215. — Duplex Steam Pump And Receiver Arranged For Automatic Return, To 
Boiler, Of Condensate From Heating Apparatus. 



principal types: (1) The combined pump and receiver (Fig. 215). 
(2) The return trap (Fig. 216). With both classes of apparatus 
the hot water or condensate which returns from the radiators 
and heating coils is collected in a receiving tank. By the 
first method, however, a direct-acting steam pump is auto- 
matically operated to discharge the water from the receiving- 
tank into the boiler. By the second method the water is 
dumped directly from the receiving-tank into the boiler. 



Sec. 236] 



BOILER-FEEDING APPARATUS 



203 



Explanation. — With The Combined Pump And Receiver (Fig. 215) 
the condensate from the heating apparatus enters the receiver through the 
inlet nozzle. When the body of water accumulates until its surface 
stands at about half the height of the receiver it buoys up the bucket- 
float B. Steam is thereby admitted to the pump through the valve V, 
the stem of which is connected to the float-lever at F. As the water- 
level in the receiver is lowered by the action of the pump the opening 



Pipe Connection Between 
Upper And Lowzr Receiver-^ 



T*$& 



Steam Supply 
For Discharging 
Trap-^ 



Upper 
Trap, 




Pipe Connection \ 
Between Receiver i 
And Trap 



Lower Rzczwer-- 
ower Trap 



Mate-Up-'' 
Water , 
Connection 



Fig. 216.— Bundy Traps Arranged For Return To Boiler Of Condensate From Heating 

Apparatus. 

through the steam valve V is gradually diminished, due to the depression 
of the float. The speed of the pump is thus regulated according to the 
quantity of water flowing into the receiver. The water which is required 
to make up for loss of steam or condensate from the system, due to leak- 
age or other cause, is admitted at M. 

With The Return-Trap Method (Fig. 216) the condensate from the 
heating system collects in the lower receiver, R h and flows thence into 
the bowl, B h of the lower trap T\. When sufficient water has accumu- 
lated in the bowl B x to cause it to tilt (Sees. 487 and 488) steam at boiler 



204 



STEAM POWER PLANT AUXILIARIES 



[Div.6 



pressure enters through the pipe Pi and forces the water into the upper 
receiver, R 2 , whence it flows into the bowl B 2 of the upper trap T 2 . 
This trap is located 3 ft. or more above the normal water-level in the 
boiler. When the bowl B 2 tilts under the weight of the accumulated 
water, steam at boiler pressure enters through the pipe P 2 . The pres- 
sure in the trap and in the boiler is thus equalized. Due to its static 
head of 3 ft. or more, the water in the bowl B 2 flows, by gravity, into the 
boiler through the feed-pipe F. When empty, the bowl tilts back to its 
filling position. 

237. The Duplex Boiler-Feeder (Figs. 217 and 218) operates 
similarly to a return trap system but has larger capacity. 



Equalizing Pipes- 




Wafer Connections 

On Reverse End—' \ \J^ : --Live Steam Inlet 
Exhaust Stzam 'Outlzi' 

Fig. 217. — Farnsworth Duplex Boiler Feeder. 

This feeder is recommended by its manufacturers for boiler- 
feeding in non-condensing plants where water from the mains 

Cho<mbzr-% ; -Chamber-fr 

3H 



C3C^..,.....CTl^.11 






>:■.'■• 




Fig. 218. — Showing Installation Of Duplex Boiler Feeder In Connection With Closed 
Heater In Non-Condensing Plant. 

is fed to the boiler through some sort of feed-water heater. 
It depends for its operation on a water supply under sufficient 
pressure to flow to the top of the boiler. 



Sec. 238] BOILER-FEEDING APPARATUS 205 

Explanation. — The feeder shown in Fig. 217 is located above the 
boiler. The tank consists of two equal compartments A and B separated 
by a central wall. It is pivoted below its center of gravity so that it 
may oscillate a few degrees in either direction. A system of valves is 
arranged in the pivot so that whichever compartment is down is allowed 
to drain into the boiler. Boiler pressure is admitted at the top of the 
compartment to make this possible. Meanwhile, the raised compart- 
ment fills with water from the feed line. Whenever the weight of water 
in the upper compartment is sufficiently greater than that in the lower, 
the tank tilts and the process in the two compartments is reversed. 

238. The Relative Merits Of Pumps And Steam Traps For 
Boiler Feeding are as follows: {Power Plant Engineering, 
Dec. 1, 1920). Where direct-return steam traps can be used to 
feed a boiler or boilers, they usually provide a more economical 
method than do steam pumps; this all depends, however, on 
the conditions in the plant. 

Explanation. — Where returns from a heating system are fed into a 
boiler, unless the boiler is low enough so that the returns can feed by 
gravity to a trap located 4 or 5 ft. above water level in the boiler, it is 
necessary to use two traps; one to force the water up to the trap above 
the boiler by means of boiler pressure steam; the other a direct return 
trap to dump water into the boiler. In such a case, the cost of a trap 
installation is, of course, higher than that for a pump, which can force 
the water directly into the boiler without rehandling. 

Where the feed water of the boiler is not made up entirely of condensed 
steam, and the load is variable so that the amount of feed must be 
varied, the speed of the feed pump can be controlled more easily than a 
trap. The trap simply dumps into the boiler whatever water comes to it, 
and, of course, the rate of flow into the trap could be regulated by the 
valve in the supply fine. Where cold water is used for feed and has to 
be heated, it is difficult to arrange the system so as to feed through a trap, 
as either the feed-water heater must be located above the trap or a lifting 
trap be employed to take water up to the direct-return trap. The chief 
argument for the pump is convenience and flexibility, and adaptability 
to all conditions. 

QUESTIONS ON DIVISION 6 

1. Name the three principal kinds of devices used in boiler-feeding. 

2. What is the chief use of injectors in stationary power-plants? Under what con- 
dition has it an economic advantage over other kinds of feeders? 

3. What is a mechanically-driven boiler-feed pump? A motor-driven boiler-feed pump? 

4. Why is mechanical drive ordinarily more efficient than electric drive for a boiler- 
feed pump? Demonstrate with an example. 

5. What is a steam-driven boiler-feed pump? What operating feature of a pump of 
this type gives it a distinct advantage over power pumps? 



206 STEAM POWER PLANT AUXILIARIES [Div. 6 

6. What is the function of a governor on a direct-acting steam pump in boiler-feed 
service? 

7. Describe the operation of a diaphragm type of pump governor. 

8. What factors mainly decide the type of boiler-feed pump that will best subserve 
the economy of a power plant? 

9. Why are centrifugal pumps generally preferable to reciprocating pumps for feeding 
boilers of installations of over 500 horsepower? 

10. What is the average steam consumption of steam-turbine-operated boiler-feed 
pumps in plants of medium capacity? What is the average mechanical efficiency of 
these pumps? 

11. Why are power-pumps better adapted than steam pumps for boiler-feeding in 
non-condensing power plants which are unequipped with heating systems? 

12. Why will downward fluctuations of the load on a boiler plant impair the economy 
of a mechanically-driven feed-pump in a greater ratio than in the case of a steam- 
driven feed-pump? Demonstrate with an example. 

13. Why should both steam-pumps and power-pumps be included in the regular 
boiler-feed equipment of a non-condensing plant which is provided with an extensive 
heating system? 

14. Describe an automatic pumping system for feeding a boiler with the returns from 
a heating system. 

15. Describe a return-trap system of boiler feeding. 

16. Explain the operation of a Farnsworth Duplex Boiler Feeder. 

17. What are the advantages and disadvantages for return traps as compared to 
pumps for boiler feeding? 

18. About what per cent, of the total coal is used indirectly by a boiler feed pump 
in a well-designed and operated plant? 

19. About what per cent, of the exhaust of a non-condensing engine is necessary to 
heat the feed water? 

20. What is meant by maintaining an exhaust-steam "heat-balance" in a power 
plant? Describe equipment for maintaining such a heat-balance automatically. 

21. What is the disadvantage of constant-speed motors for feed-pump drives? What 
kind of motor is free from this disadvantage? 

22. What are two disadvantages of centrifugal pumps as stand-by boiler-feeding 
equipment? 

23. Why should reciprocating power pumps be fitted with relief valves under some 
conditions? 

24. What are the two general methods of estimating feed water requirements? 
Explain each. 

25. What is meant by "forcing" a boiler? How much may one be forced? With 
what results? Under what conditions? 

26. Name four common troubles of pump governors and give their remedies. 

PROBLEMS ON DIVISION 6 

1. A set of boilers has a total rating of 600 boiler h.p. If it is desired to have a 
pump capacity of 50 lb. of water per hr. per boiler h.p., what should be the rating of 
the pump in gallons per hour. If it is later decided to force the boilers 225 per cent, 
at peak load, what capacity should the pump then have if it is to have the same per 
cent, excess capacity as before? 

2. The main engine of a power plant has a duty of 150 million ft. lb. per 1,000 lb. 
of steam and develops 500 h.p. If the auxiliaries require 10 per cent, as much steam 
as the main engine and it is desired to have a feed pump capacity 50 per cent, in excess 
of normal requirements, how many gallons per hour must the pump deliver? 



DIVISION 7 
FEED -WATER HEATERS 

239. The Reasons That Feed-Water Heaters Should Be 
Used, Fig. 218A, are these: 

(1) If cold water is fed into a boiler, additional fuel must be 
burned to raise its temperature almost to the boiling point. 
This represents a costly waste of fuel, inasmuch as in practically 
every plant either exhaust steam or hot flue gases or both, which 
would otherwise be dissipated into the atmosphere and lost, can 
be used for feed-water heating. 

(2) The steel plates of a boiler which is in operation are very 
hot. If cold water is fed into it, certain parts of the boiler shell 
may thereby be cooled excessively. Thus high stresses will be 
produced due to unequal expansion of the shell. The plates are 
strained as are also the riveted joints. Leakage at the joints and 
decreased life of the boiler may result. 

(3) When cold water is pumped into boilers, it may contain 
impurities, which tend to form scale on the inside of the boilers 
when it becomes hot. This scale not alone interferes with the 
rate of transmission of heat from the fire to the water, but it also 
may permit certain parts of the shell to become excessively hot 
because the water in the boiler is prevented by the scale, from 
contacting intimately with the shell. Blistering and short 
boiler life may result. But if the water is first heated to at least 
200 deg. fahr. before being forced into the boiler, many of these 
impurities may be thereby precipitated in an external chamber, 
from which they can be removed readily. Thus they are prevented 
from entering the boiler. 

240. In A Non-Condensing Plant Eighty Per Cent. Of The 
Energy In The Live Steam Is Wasted In The Exhaust. — That 
is, the amount of heat remaining in the exhaust steam from a 
non-condensing engine is about 80 per cent, of the original 
heat imparted in the boiler to the steam. The truth of this 
may be shown thus: 

207 



208 



STEAM POWER PLANT AUXILIARIES 



[Div. 7 




Sec. 240] FEED-WATER HEATERS 209 

Example. — Consider a medium-capacity, well-maintained non-con- 
densing plant operating at 150 lb. per sq. in. boiler pressure. Such a 
plant should develop an indicated horsepower hour (i.h.p. hr.) on 25 lb. 
of steam. That is, its water rate would be 25 lb. of steam per i.h.p. hr. 
Assume that the cold feed water has a temperature of 50 deg. fahr. 
Hence, we are interested only in the heat which must be added to this 
cold feed water to raise it to the temperature of steam at 150 lb. per sq. in. 

From a steam table it is found that the heat which must be added to 
1 lb. of water at 50 deg. fahr. to convert it into steam at 150 lb. pressure 
is 1,177 B.t.u. Hence, on this basis the 25 lb. of steam which is required 
by the engine to produce 1 h.p. hr. represents: 25 X 1,177 = 29,425 
B.t.u. Now, from a conversion table, it is found that 1 h.p. hr. is equal 
to 2,545 B.t.u. Therefore, out of the 29,425 B.t.u. imparted to each 
pound of steam, only 2,545 B.t.u. is converted into useful work in the 
production of 1 h.p.hr. Thus there must be in the exhaust steam from 
the engine (disregarding radiation) : 29,425 — 2,545 = 26,880 B.t.u. per 
i.h.p. hr. The percentage of heat converted into work on the engine 
piston must, then, be: 2,545 -=- 29,425 = 0.087 = 8.7 per cent. 

If the radiation losses are assumed to be 10 per cent, (of the heat in 
the exhaust steam) which is a fair average value, the available heat per 
indicated h.p. hr. would be: 26,880 - 2,688 = 24,192 B.t.u. The 
percentage of the total heat received by the engine which is lost in radi- 
ation is: 2,688 -s- 29,425 = 0.091 =9.1 per cent. Hence the percentage 
(of the original heat which was in each pound of steam) that is now 
available in the exhaust is: 24,192 ~ 29,425 = 0.822 = 82.2 per cent. 
Note, then, that about 82 per cent, of the original heat is available in 
the exhaust steam from the engine. Thus, summarizing, the percen- 
tages of the heat units delivered to the engine cylinder in the live steam 
are either expended or available thus: 

Heat expended as work on engine piston 8.7 per cent. 

Heat lost in radiation 9.1 per cent. 

Heat available in exhaust steam 82 . 2 per cent. 

Total 100 . per cent. 

Where larger engines and turbines operating condensing and with 
superheat are used, a greater proportion of the heat is realized in useful 
work. A non-condensing prime mover discharges its exhaust steam into 
the atmosphere at 212 deg. fahr. A condensing prime mover discharges 
its exhaust steam into its condenser at a temperature of about 100 deg. 
fahr. or lower, depending on the vacuum maintained. Even with the 
most-efficient, condensing, steam-power-plant equipment, where the 
water rate is as low as 10 lb. of steam per h.p. hr., about 75 per cent, of 
the heat is discharged with the engine exhaust and is, for all practical 
purposes, lost. 

14 



210 



STEAM POWER PLANT AUXILIARIES 



[Div. 7 



241. The Two General Types Of Feed-Water Heating 
Equipment are : (A) Exhaust steam feed-water heaters (which 
are treated in this Division) which are devices which use the 
exhaust steam for raising the temperature of the feed- water. 

(B) Economizers (Div. 8) which 
are devices which use, for heat- 
ing the feed water, the hot flue 
gases after they are discharged 
from the boiler-furnace. 

242. Exhaust Steam Feed- 
Water Heaters are of many 
types but may be classified into 
two general divisions: (1) The 
open heater, Fig. 219. (2) The 
closed heater, Figs. 220 and 221. 
By an open heater is meant one 
in which the exhaust steam is 
permitted to contact directly in a suitable chamber with 
the cold water which is to be heated. Thus part of the 
exhaust steam is condensed in raising the temperature of 
the cold water and is used as part of the feed water. With 




^To Feed Pump '-Hot Wafer 



Fig 



219. — Diagram Of Open Feed- 
Water Heater. 



Safety Valve 
Feed Outlet^ Connection- -j> 




Fig. 220.— "Blake-Knowles" Water-Tube Type Of Closed Exhaust-Steam Feed- 
Water Heater. (The water passes through each of the six tube nests in turn, 
thus traversing the heater six times. The steam passes three times through the 
heater.) 



an open heater, the temperature of the feed water can — 
assuming that sufficient exhaust steam is available, and there 
usually is, be raised to a temperature of 210 to 212 deg. By 



Sec. 243] 



FEED-WATER HEATERS 



211 



a closed heater is meant one in which the steam does not contact 
with the cold water but in which the heat from the exhaust 
is imparted to the water through the walls of tubes. Thus 
in the closed type, the water to be heated and the exhaust 
steam for heating it are confined to separate chambers. 



.-Shell 



Water From FeeolPump-s 
Tubes-. 




•OufkthBoile, 



Fig. 221.— Diagram Of Clo: 
Feed- Water Heater. 



ed 



Note. — The Closed Heater Must 
Be Used Where The Boiler Feed- 
Water Must Be Maintained Abso- 
lutely Free From Oil. An oil 
separator, which extracts practically all 
of the oil, always forms a part of open 
heater equipments, but these separators 
cannot always be relied upon to extract 
all of the oil from the exhaust steam. 



243. Economies Accruing Due 
To The Use Of Feed-Water 
Heaters are very pronounced. In 
the average plant a saving of from 
11 to 14 per cent, in fuel may be 
expected due to the installation of 
a heater. There is usually sufficient 
exhaust steam (see Sec. 209) which would otherwise be wasted, 
available to heat the feed water. All of the heat which can 
be imparted to the feed water before it is pumped into the 
boiler represents that much saving in fuel. A temperature of 
212 deg. fahr. is the highest to which water can be raised (at 
atmospheric pressure) without its being converted into steam. 
It follows that every effort should be made to utilize exhaust 
steam to raise the feed water to 212 deg. fahr. While a tem- 
perature of 212 deg. fahr. may not be feasible in every case, it 
is usually possible to attain a feed-water temperature of 210 
or 211 deg. fahr. Because a higher feed-water temperature 
can be obtained with an open heater than with a closed one, 
the open type is somewhat more economical. Every steam- 
power plant should have a feed-water heater. 

Note. — The Following Rules For Estimating The Approximate 
Fuel Saving Due To Preheating Feed- Water are often useful: 
(1) For every 11 deg. fahr. which is added to the temperature of the feed- 
water with exhaust steam there results a saving of about 1 per cent, of the 
fuel which would otherwise be required. (2) For a given consumption of 



212 STEAM POWER PLANT AUXILIARIES [Dtv. 7 

fuel, the evaporative capacity of a boiler is increased by approximately 1 
per cent, for each 11 deg. fahr. increase of the feed-water temperature. 

Explanation. — Suppose the temperature of the feed-water is 60 deg. 
fahr., and the boiler pressure is 120 lb. per sq. in., gage. According to 
the steam tables, found in any engineering handbook, the total heat, 
above 32 deg. fahr., of steam at 120 lb. per sq. in., gage, is 1191.6 B.t.u. 
per pound. Therefore, the total heat that must be supplied to each pound 
of the feed-water is [1191.6 - (60 - 32)] = 1163.6 B.t.u. If, now, the 
feed-water temperature is raised to 71 deg. fahr. by waste heat, the 
saving = (71 — 60) =11 B.t.u. per pound. Then the per cent, saving = 
11 -J- 1163.6 = 0.009,5 or roughly 1 per cent, of the total heat supplied to the 
steam. 

Example. — A power plant, in which the boilers develop 1,000 boiler 
h.p. with feed water at 100 deg. fahr., is furnished with a heater which 
supplies the feed water at 210 deg. fahr. What additional boiler horse- 
power is thus realized? 

Solution. — By Sec. 243 the evaporative capacity of the boilers is 
increased approximately 1 per cent, for each 11 deg. fahr. increase of the 
feed-water temperature. Hence, the power of the boilers is increased 

po-_ioo + 10Q j x 1000 = 100 hp 

244. The Saving Of Heat Which Results From Preheating 
Boiler Feed-Water with exhaust steam that would otherwise 
be wasted may be computed by the following formula: 

(76) H f = gZ ^T ^32) 10 ° (per cent - } 

Wherein Hf = the saving, in per cent, of the heat-content of 
the fuel. Tfi = the temperature of the feed-water, in degrees 
Fahrenheit, before preheating. T /2 = the temperature of 
the feed-water, in degrees Fahrenheit, after preheating. 
H = the total heat in the steam which is generated in the 
boiler, in British thermal units per pound. 

Note. — The specific heat of water varies somewhat with the tempera- 
ture (see the author's Practical Heat). In the compilation of For. 
(76), however, the specific heat of the feed-water is assumed to have a 
constant value of 1.0 B.t.u. per lb. for all temperatures. Computations 
based upon this assumption are correct within 1 per cent., which is suf- 
ficiently accurate for all practical purposes. 

Example. — A boiler generates steam at a pressure of 100 lb. per sq. 
in., gage. The water which is fed to the boiler is preheated, with exhaust 
steam, from 80 deg. fahr. to 210 deg. fahr. What saving of heat results 
from thus utilizing the exhaust steam? 



Sec. 245] 



FEED-WATER HEATERS 



213 



Solution. — As given in a table of the properties of saturated steam, 
the total heat in steam at 100 lb. per sq. in., gage, is 1188 B.t.u. per lb. 
Hence, by For. (76), the saving = H f = { (T /2 - T fl )/[H - (T fl - 
32)]} 100 = {(210 - 80) h- [1188 - (80 - 32)1} X 100 = 11.4 per cent. 

245. The Percentage Of Fuel Saving Due To Feed-Water 
Heating May Be Computed Graphically (Fig. 222) for satu- 
rated or superheated steam. Points A and C, for instance, 
are found corresponding to initial and final feed-water tem- 
peratures. A vertical line from A is traced until it intersects 
an oblique line from C at B. A point D is then found on the 



5a fur af ed Steam- - 
Temp.Vfater Leaving Heater- 




C 150 140 A 60 
240 200 160 120- 60 40 
Initial Feed-Water. Temperature Deg.Faht: 

Fig. 222. — Graph Showing Percentage Of Fuel Saved By Heating Feed Water. 

scale at the upper left corresponding to the steam gage pres- 
sure. A vertical line from D is traced to its intersection with 
a graph for saturated steam at E or some degree of superheat 
at F. Lines from F and E are traced horizontally to the line 
AB and then obliquely until they intersect a horizontal line 
from B at G and H. The saving in each case may be read 
from the per cent, scale. 

Example. — In the case selected in Fig. 222 the initial temperature 
was, A, 110 deg. fahr. The water left the heater at C, 210 deg. fahr. 
The gage pressure was, D, 160 lb. per sq. in. For saturated steam, the 
saving was, G, 9.1 per cent. For 100 deg. fahr. superheat the saving 
was, H, 8.7 per cent. 

246. The Net Monetary Saving Which Results From 
Preheating Boiler Feed -Water With Exhaust Steam that 
would otherwise be wasted must be computed upon a basis 



214 STEAM POWER PLANT AUXILIARIES [Div. 7 

of the interest on the investment in heating apparatus and the 
annual cost of depreciation, attendance and maintenance, 
taken in conjunction with the annual heat-saving effected, 
which may be computed by using For. (76). 



Example. — The coal-consumption of a battery of boilers which receive 
feed-water at a temperature of 110 deg. fahr. is 3 tons per day. It is 
estimated that by utilizing a quantity of exhaust steam which is now 
going to waste, the feed-water may be preheated to 212 deg. fahr. The 
average steam-pressure is 110 lb. per sq. in., gage. The coal costs $3 
per ton. The plant operates 310 days per year. The cost of the feed- 
water-heating apparatus and its installation will be $300. The rate of 
interest on the investment is 6 per cent, per annum. The assumed rate 
of depreciation is 5.0 per cent, per annum. The cost of maintaining and 
operating the apparatus is presumed to be $5 per month. What will be 
the probable annual net saving? 

Solution. — As given in a table of the properties of saturated steam, 
the total heat in steam at a pressure of 110 lb. per sq. in., gage, is approxi- 
mately 1,190 B.t.u. per lb. Hence, by For. (76), the probable thermal 
saving = H f = {(T /2 - T fl )/[H - (T/i - 32)]} 100 = {[(212 - 110) -r- 
[1,190 - (110 - 32)]} X 100 = 9.17 per cent The present annual cost 
of the coal supply = 3 X 3 X 310 = $2,790. Therefore, the probable re- 
duction in the annual coal bill, due to utilizing the available exhaust 
steam = (2,790 X 9.17) -r- 100 = $255.84. The interest on the invest- 
ment = (300 X 6) -7- 100 = $18. The annual cost of depreciation = 
(300 X 5.0) -7- 100 = $15.00. The annual cost of maintenance and oper- 
ation = (12 X 5) = $60. Hence, the approximate net annual saving will 
be 255.84 - (18 + 15 + 60) = $162.84. 

In other words the feed-water-heating equipment will pay for itself in 
about 2 yr. If the heater were installed in a plant where it would not 
be necessary to employ additional labor to maintain it, it would pay for 
itself in about 1% yr. 

Note. — Op All Boiler Room Accessories, Feed-Water Heaters 
Are, Probably, The Most Effective Savers Of Coal. (From the 
American Correspondence School.) With condensing engines, the 
condensate-pump discharges from the condenser into the hot well. Then 
the water is drawn from the hot well as boiler feed at a temperature of 
100 deg. to 140 deg. F. This, however, if the boiler pressure is over 
100 lb. per sq. in., is not a sufficiently-high temperature for the best 
economy. Feed water at this temperature should be passed through a 
feed-water heater. With non-condensing engines it is, from a standpoint 
of economics absolutely necessary that in some way the feed water be 
heated by the exhaust steam in a feed-water heater or by the waste gases 
from the chimney in an economizer. 



Sec. 247] 



FEED-WATER HEATERS 



215 



247. Exhaust-Steam Feed-Water Heaters May Be Classi- 
fied With Respect To Their Relation To Other Plant Equip- 
ment (see also Sec. 242), as hereinafter explained, as primary 
and secondary heaters. They may be classified with respect to 
the steam pressure used as atmospheric, vacuum and pressure 
heaters. 

248. Table Showing Classification Of Representative 
American Feed-Water Heaters (partly from Gebhardt). 



Exhaust 
steam 


Open 
Atmospheric 


Bonar Moffat 
Blake-Knowles Reliance 
Cochrane Sims 
Cookson Stillwell 
Elliot Webster 
Hoppes 


T3 
O 


Vacuum, pressure, 
or atmospheric 
(water tube). 


American Ross 
Griscom Russel Standard 
Gaubert Wainwright 
National Wheeler 




Vacuum, pressure, 
or atmospheric 
(steam tube). 


Berryman Otis 
Kelly Ross 


Live steam 


Open pressure 


Hoppes 
Baragwanath 



249. A Primary Or Vacuum Heater is a closed feed- water 
heater which is connected to the exhaust of a condensing engine 
between the engine and the condenser. The conditions favor- 
able to the installation of a primary heater exist where the 
supply of exhaust steam from the auxiliaries in a condensing 
plant is insufficient for properly heating the feed water. In 
such cases (Fig. 223) the feed-water can first be heated in the 
primary heater, with steam exhausting from the engine and 
then be passed through a secondary heater (Sec. 250) which is 
supplied with exhaust steam, at atmospheric pressure or above, 
from the auxiliaries. The primary heater is under about the 
same vacuum as the condenser. If the condenser maintains 
a vacuum of 26 in., the temperature of the discharge from the 



216 



STEAM POWER PLANT AUXILIARIES 



[Div. 7 



primary heater will probably not exceed 118 deg. fahr. The 
primary heater also acts as a supplementary surface condenser 
in which the feed-water acts as condensing water. 

Note. — A Primary Heater Is Especially Useful Where A Jet 
Condenser Is Used And The Condenser Water Is Unsuited For 
Boiler Feed. When this is true, fresh water must be used as boiler 
feed and is usually supplied at much below hot-well temperatures. For 
instance, if an average hot-well temperature is 100 deg. fahr. and the 



flh 



LiYZ-Stzam P/pe ; 




-Siphon 
Cono/enser 



St 0am Supply 



Fig. 223. — Showing Method Of Installing Primary And Secondary Feed- Water Heaters. 

water supply temperature is 60 deg. fahr., additional heater capacity is 
necessary to raise the water the difference of 40 deg. fahr. The same 
necessity for a primary heater exists when a high vacuum is obtained 
with a surface condenser. The condensate may then be cooled to 60 
deg. fahr. or a lower temperature. 

250. An Atmospheric Heater is an open or closed feed- 
water heater (Fig. 224) which utilizes the exhaust from non- 
condensing engines or auxiliaries. The pressure on these heaters 
is equal to the back-pressure on the engines which supply 
the exhaust steam. Where auxiliaries supply the exhaust, 
this pressure is usually controlled by a back pressure valve, at 



Sec. 251] 



FEED-WATER HEATERS 



217 



a few pounds above atmospheric. Where the exhaust from 
the heater is used in a vacuum heating system, the pressure 
may be a few inches mercury column below atmospheric. The 
maximum feed-water temperatures obtainable in atmospheric 
heaters are about 200 deg. fahr. in closed heaters and 210 deg. 
fahr. in open heaters. 




'-Boiler Feed Pump '-Through Type Of Exhaust-Steam Feed-Water Heater 

Fig. 224. — Showing Piping Arrangement Of Stilwell Through Type Of Exhaust-Steam 
Feed-Water Heater In Non-Condensing Plant. 

251. Both Vacuum And Atmospheric Heaters May Be 
Used In Condensing Plants (Fig. 223). — The feed- water is 
first forced by the feed pumps through the vacuum heater, in 
which it absorbs whatever heat may be abstracted from the 
exhaust steam coming from the main engine. The feed-water 
then passes through the atmospheric heater and on to the 
boilers. Exhaust steam from the pumps or from any other 
source, which it may be inconvenient or unprofitable to con- 
dense, is piped to the atmospheric heater. When an atmos- 
pheric heater is connected in this way it is commonly called a 
secondary heater as distinguished from a primary or vacuum 
heater. 

Note. — The Secondary Heater Mat Be Of The Open Or Closed 
Type. When, however, it is of the open t} r pe, the feed water must flow 
by gravity — or be forced by a separate pump — through the primary 



218 



STEAM POWER PLANT AUXILIARIES 



[Div. 7 



heater. It is usual therefore to select secondary heaters of the closed 
type. The primary heater is always of the closed type. 

252. Installation Of Primary And Secondary Feed -Water 
Heaters, To Be Operated Alternately (Fig. 225), may be advis- 
able for condensing plants in which the quantity of exhaust 
steam from the auxiliaries is, ordinarily, sufficient for feed- 
water heating, but where the condenser auxiliaries are occa- 



: To hkatingr Or 
"Dryfngr System 




■Discharge': 



Feed-Sump Suction-' K Steam-Supplu Pipe-' K AirPump 



feed-Pump Suction-' ^Steam-Supply Pipe- '^AirPump 

Fig. 225. — Installation Of Primary And Secondary Heaters For Alternate Operation. 

sionally inoperative on account of the main engine being 
required to exhaust to the atmosphere. With such installa- 
tions the primary heater can be used alone at such times as the 
main engine is running non-condensing, while the secondary 
heater can be used alone when the operation is condensing. 

253. The Back-Pressure On An Engine May Not Be 
Increased By Installing A Feed-Water Heater in the exhaust 
line. This is, with closed heaters, due to the fact that the 
shell of the heater, in the case of a water-tube heater, or the 



Sec. 253] 



FEED-WATER HEATERS 



219 



nest of tubes, in the case of a steam-tube heater, is, usually, 
of much greater cross-sectional area than the exhaust pipe. 
Also, the partial condensation of the exhaust steam, due to 
absorption of heat therefrom by the feed-water, tends to 



To Atmosphere 
Mult/port /a/ye 



Exhaust To Heating And Drying Coi/S\ 
/■Co/a' Water Supply \ 




s/WW/////////W//<\ 



^Lighting Circuits \ 
lighting Company's Street Circuit-' 



~-Turbine-Dr/ren Induction 
Generator 



Fig. 226. — Cochrane Open Induction Heater, H, Equipped With Automatic Thermo- 
static Valve, Used For Exhaust Steam-Heating System. 
(In the ordinary power plant which uses exhaust steam-heating, the power and heating 
requirements rarely balance. Some of the time, perhaps half the year, steam is wasted 
to atmosphere. On the other hand, central-station energy may be used when more 
power is required than the heating system will generate as a by-product. Or, an auto- 
matic heat balance for such conditions may be provided by the arrangement shown 
above. The back-pressure turbine, T, exhausts into a steam-stack heater, H, and to 
the heating system, S. The generator, G, supplies energy to the local circuit, which is 
also connected to the central-station company's street mains through a meter. 

A thermostat, responsive to the temperature of the water in the heater, governs the 
admission of steam to the turbine (subject, of course, to an automatic speed limit). 
When the power requirements are greater than the heat requirements, central-station 
energy is taken through the meter. If at times more heat than power is required, steam 
can be by-passed automatically or power can be sold back to the electric company. 
The conversion of heat to mechanical power and building heating is, with this arrange- 
ment, practically 100 per cent, perfect. No heat is wasted to atmosphere or to con- 
denser circulating water.) 



prevent back-pressure. An open induction heater (Sec. 254) 
with an extra-large oil-separator may be used on a non-con- 
densing engine exhaust without increasing the back-pressure 



220 



STEAM POWER PLANT AUXILIARIES 



[Div. 7 



more than J^ lb. per sq. in. Where the engine exhaust is 
used for feed-water heating only, an open heater arranged 
thus and properly vented and managed will heat the feed 
water to within about 2 deg. fahr. of the exhaust steam 
temperature. Some open heater manufacturers claim that 
an open heater need not cause any additional back-pressure. 

Note. — A Back-Pressure Valve Increases The Effectiveness Of 
A Feed-Water Heater and acts also as a safety valve for the heater. 
It should be a reliable easy-moving valve of large port opening. When 
an induction heater and a heating system are supplied with steam from 
the same exhaust line (Fig. 226), a back-pressure valve is necessary to 
insure proper distribution of the steam to all the heating equipment. 
A back-pressure valve decreases the power developed by an engine about 
2)4 per cent, for each pound of back pressure. The cost of the decreased 
engine efficiency due to back pressure carried for a heating system is 
usually much less than the cost of the live steam which would be required 
for heating if the back pressure were not maintained. The decrease in 
engine power due to a back pressure may be made up by carrying 2 to 5 
lb. per sq. in. greater boiler pressure for each pound of back pressure — 
which will of course require the burning of a slightly greater amount 
of coal. 



Exho/usf To Atmosphere- 



Exhaust 
From 
Engine 



254. Exhaust-Steam Feed -Water Heaters May Be Classi- 
fied According To Their Piping Arrangements as: (1) Induced 

or draw heaters (Figs. 227, 228, 
229, 230, 231, 232), which re- 
ceive no more exhaust steam 
from the available supply than 
the water will entirely condense. 
(2) Through or thoroughfare heat- 
ers (Figs. 233 and 234), which 
receive all of the available 
supply of exhaust steam. With 
the first arrangement, complete 
condensation of the steam 
which passes into the heater 
induces a continual flow thereto 
through a branch from the main 
exhaust pipe. If the quantity 
of steam exhausted by the engine is greater than that which 
can be condensed in the heater, the excess, with the first 




Drain-* 



\\\\\\\\A\\)A\\\\\V. 

Water Outlet-'' ( I 
Drain Connection^' 



Fig. 227. — Horizontal Closed Heater 
Piped For Service On The Induction 
Principle. 



Sec. 254] 



FEED-WATER HEATERS 



221 



arrangement, may go directly from the engine to the atmos- 
phere, or to a heating system or condenser. With the second 



-■Risers To Heating System. 



■Exhaust Head 




f^^^^^^^^^^^^^^^^^^: 



Exhaust Main- 



Fig. 228. — Hoppes Horizontal Exhaust-Steam Feed-Water Heater Installed For 
Induction Operation With Gravity Heating System. 



ixhaust Steam To Atmosphere 

W-Back Pressure Valve H'9 h Pr^ rz 

a Steam Ma/n—-± 
til ~ 




SSSJ? 



Fig. 229. — Typical Installation Of Open Feed-Water Heater Pipes For Induction 
Operation In Connection With Vacuum Heating Plant. 

arrangement, if more steam is received than can be condensed 
in the heater, the excess passes through the heater to the 
atmosphere or heating system. 



222 



STEAM POWER PLANT AUXILIARIES 



[Div. 7 



Note. — If The Quantity Of Exhaust Steam Available For Heat- 
ing The Feed- Water Is Excessive, the open heater should (Fig. 228) 
be arranged for induction service. If the surplus exhaust steam from 




-To Heating System Or 
To Atmosphere 



^Impulse Of Steam Current 

Fig. 230. — Impulse Of Steam Current Di- 
rected Toward Induction Heater. 




^W\v\\\^^ 



Fig. 231. — Impulse Of Steam Current At 
Right Angles To Induction Heater. 



■^■■Back-Pressure Valye ^-Pece'ver Separators 




Wm??77777m77^ 



Fig. 232. — Induction-Type Open Feed-Water Heater, H, Installed In Connection 
With Reciprocating Engine, E, And Mixed-Flow Turbine, T. (Exhaust from recipro- 
cating-engine, feed-pump turbine, F, and auxiliary turbine, A, piped to feed-water 
heater, heating system and mixed-flow turbine.) 

an induction heater is used for heating or drying purposes, and the result- 
ing condensate is afterwards returned to the heater, the surplus steam 



Sec. 255] 



FEED-WATER HEATERS 



223 



should be passed through an independent oil separator. With induction 
operation the surplus steam will pass on in a much drier condition than 
if it had gone through the heater. If the condensate from a closed heater 



Air Outlet--. 
Exhaust Outlet-, 
To Atmosphere 

Feed Pump 
Exhaust Pipe-, 




'Boiler Feed Pump Condensate Pump-'' Surface- 

Condenser 

Fig. 233. — Showing Piping Arrangement Of Stilwell Through-Type Of Exhaust-Steam 
Feed- Water Heater In Condensing Plant. 



To Boilers, 



-To Heating System Or Atmosphere 




Fig. 234. — Equipment Of Closed Feed- Water Heater Installed For Service On The 
Thoroughfare Principle. 

is to be returned to an open heater, the inlet to the closed heater should 
be fitted with an oil separator. 

255. The Piping Of An Induction Heater should be so 
arranged, when possible, that the direct impulse of the exhaust- 
steam current (Fig. 230) is toward the heater, rather than 



224 



STEAM POWER PLANT AUXILIARIES 



[Div. 7 



Exhaust Outlet- 
V 



Sfee/ 5/?e// N 



toward the atmosphere or heating system (Fig. 231). The 
object of this is to insure delivery to the heater of as much 
steam as it can accommodate. With the impulse of the 
steam current at right angles to the heater (Fig. 231) the 
heater might receive a scanty or starved supply. 

256. The Temperature Of The Exhaust Steam Entering A 
Feed-Water Heater depends upon the back-pressure. If the 
steam in excess of that which is condensed in the heater is 

discharged directly to the 
atmosphere, then the back 
pressure is, ordinarily, at- 
mospheric pressure. 
Hence, in such cases the 
temperature is about 212 
deg. fahr. But if the ex- 
cess of steam is used in a 
heating system, the back 
pressure may range from 
atmospheric up to about 5 
lb. per sq. in. In the lat- 
ter case the temperature 
would be about 227 deg. 
fahr. 

257. Open Exhaust- 
Steam Feed -Water Heat- 
ers Are Generally Designed 
To Perform A Four-Fold 
Function as follows : (1) To 
remove the oil from the ex- 
haust steam which supplies 

mmmmmMrmffitiMffifr the heaL This is accom - 

Fig. 235.— The Moffat Open Exhaust-Steam pHshed by means of an 
Feed- Water Heater And Purifier. oil-Separating device (Fig. 

236) which (Fig. 235) usually forms an integral part of 
ihe heating apparatus. (2) To bring the exhaust steam and 
feed-water into intimate contact. The heating effectiveness of 
the apparatus depends principally upon the thoroughness 
with which this detail of its operation is fulfilled. (3) To 
purify the mixture of feed-water and condensed exhaust steam 




Sec. 257] 



FEED-WATER HEATERS 



225 



by filtration. This may be accomplished (Fig. 235) by 
causing the heated water to percolate through chambers filled 
with filtering material. (4) To afford storage space for the 



.-External Shell 




Fig. 236.— Oil-Separating Element Of Moffat Open Exhaust-Steam Feed-Water Heater. 




") Drying Coils-. 



■-Calendering 1 Stand Pipe- ^ L. 
Pol I 5 V ' r o 



Absorption Ice 
tr-— Machine 



-Cooking Kettles 
Exhaust Turbine^ 




Exhaust From Engines, / 
Pumps Etc 

From Water Main- *■ 



■ Boiler Feed-Pump-' 
^Feed-Water Heater 
..-To Sewer Boilers- 



Fig. 237. — Diagram Showing How A Feed-Water Heater Serves As A Clearing House 
For All Available Supplies Of Exhaust Steam And Water Which Are Suitable For 
Boiler Feeding. (Light lines represent exhaust-steam piping; heavy lines, water 
piping.) 

heated and filtered water and act as a receiver for condensate 
from various sources (Fig. 237). 

Explanation. — The feed-water enters the heater (Fig. 235) through 
the pipe F. The rate of flow is controlled by the valve V, which is oper- 
15 



226 



STEAM POWER PLANT AUXILIARIES 



[Div. 7 



ated by the float H. The water rains down through the perforated 
plate R and passes successively through the filter beds Mi and M. From 
the filter bed M, the water rains down through chamber A, whence it 
percolates upward through the coke filter in chamber N, and thence 
through the strainer L into the storage chamber Y. 



Cold Wcttir 
.Supply 




\ P.ump 
Supply 



WmH, 



"FQunot'oifiQn 



Fig. 238. — Cochrane Open Induction Feed- Water Heater. 



The exhaust steam enters the heater (Fig. 235) through the nozzle E, 
and (Fig. 236) is diverted to a downward flow, through the cups C, into 
the separating chambers S. The momentum of the oil-particles precipi- 
tates them to the bottoms of these chambers. As the oil accumulates, 
it flows through the openings (R, Fig. 236) into the space surrounding 



Sec. 258] FEED-WATER HEATERS 227 

the separating chambers, and thence out through the drain-pipe W. The 
steam (Fig. 235) circulates upward through the core-pipe, K, and is 
deflected by the plate, D, through lateral openings in the core-pipe, into 
the annular chamber A. A portion of the steam is condensed by the 
water percolating through the filter bed M, another portion ascends 
through the duct T, while a considerable portion reenters the core-pipe, 
K, through the openings above the deflecting plate D. The steam which 
reenters the core-pipe is deflected into the annular chamber A x by the 
plate D\. The same events which followed the entrance of the steam 
into chamber A then ensue. Some of the steam is condensed in the filter 
bed Mi, some of it passes up through the duct 7\, while the remaining 
portion again reenters the core-pipe through the openings above deflect- 
ing plate Di. The steam ascending through the core-pipe finally en- 
counters the cold water supply as it trickles down through the rain-plate 
R. Then if the heater is operated on the through principle the uncondensed 
steam passes around the edge of the upper baffle, U, and out through 
the nozzle O. With induction operation the exhaust outlet, 0, is closed 
except for a small vent pipe leading back to the exhaust pipe (Fig. 227). 

The perforated pipes B and X have external connections, through the 
shell, to a source of water under pressure. Pipe B is provided for flush- 
ing down the coke filter. Pipe X is provided for washing the sludgy 
deposits from beneath the coke filter out through the blow-off valve. 

Note. — The Condensate From A Gravity Heating System may 
be piped, G (Fig. 235), directly to an open feed-water heater. See also 
Fig. 237. 

Note. — The same operations are performed with different construction 
by the Cochrane heater (Fig. 238). 

258. If The Carbonates Of Lime, Magnesia And Iron Are 
Dissolved In A Feed Water, they may be removed by an 
open feed-water heater. These impurities precipitate at 
temperatures below 212 deg. fahr. Hence, if this temperature 
is maintained in an open feed-water heater the impurities 
mentioned will be deposited in the heater. Thus the forma- 
tion of scale in the boilers may be largely avoided. For the 
destructive effects of scale on boiler tubes and plates see the 
author's Steam Boilers. 

259. Only Liquid Oil Can Be Removed By The Oil-Separator 
Of An Open Feed-Water Heater. — Hence, if a low grade 
of oil is used for engine-cylinder lubrication, the separation 
may not be complete. This will be due to the fact that some 
of the constituents of low grade oils vaporize at the steam 
temperature. The oil vapor will then pass into the heater 
and form an emulsion with the water. Thus a portion of the 



228 STEAM POWER PLANT AUXILIARIES [Div. 7 

oil will be delivered to the boiler. Therefore, none but a 
high grade of oil should be used for engine-cylinder lubrication 
where the exhaust from the engine is to be condensed in an 
open feed-water heater. See the Author's Steam Boilers. 

Note. — Oily Feed Water Is Very Objectionable. Oil is a very- 
poor conductor of heat. Hence, if the oil, which may be admitted to a 
boiler with the feedwater, lodges on the fire-sheets or tubes, overheating 
of the sheets or tubes may result. The overheating may then cause the 
plates or tubes to bag or bulge, thus weakening the material and inviting 
rupture. (See the author's Steam Boilers.) Hence, removal of the 
oil from the exhaust steam which is used is a very important function of 
the open feedwater heater. 

260. The Air And Carbonic Acid Gas Which Water For 
Boiler-Feed Generally Holds In Solution are largely liberated 
in an open feed-water heater at about 210 deg. fahr. If the 
separation takes place in the heater no damage will result. 
The liberated air and carbonic acid gas will pass out through 
the vent to the atmosphere. But, in the absence of an open 
heater, if the separation takes place in the boiler, the liberated 
gases will combine chemically with the material of its construc- 
tion and rapid corrosion will result. 

261. The Use Of A Feed-Water Heater Is Advisable As A 
Boiler Protective Measure Even Where No Economic Saving 
Is Apparent. — The strains in boiler plates, due to cold feed- 
water striking directly against them, are estimated (The 
Locomotive) at 8,000 to 10,000 lb. per sq. in. This in addition 
to the normal strain produced by steam pressure is quite 
enough to tax the girth seams beyond their elastic limit if the 
feed pipe discharges anywhere near them. Hence, it is not 
surprising that girth seams develop leaks and cracks in 99 
cases out of every 100 in which the feed discharges directly 
against the fire sheets. From the foregoing it is evident that 
the feed-water heater is a necessary part of the equipment of a 
power plant aside from all purely economic considerations. 

262. The Temperature To Which Feed Water May Be 
Raised By Steam In An Open Heater depends upon the 
quantity of exhaust steam available, the initial temperature 
of the feed-water, and the temperature of the exhaust steam. 
When all of the exhaust steam which is delivered to the heater 



Sec. 263] FEED-WATER HEATERS 229 

is condensed therein the final temperature of the feed water 
may be computed by the following formula: 

„-_* m T fl Wj + 0.9W S (H + 32) , , „ , , ... 

(77) T f 2 = w + 9W (degrees Fahrenheit) 

Wherein F/2 = the temperature of the water leaving the 
heater, in degrees Fahrenheit. T/i = the temperature of 
the water entering the heater, in degrees Fahrenheit. W f 
= the weight of the feed-water entering heater, in pounds per 
hour. W 5 = the weight of the exhaust steam, in pounds per 
hour. H = the total heat, above 32 deg. fahr. in the exhaust 
steam, in British thermal units per pound. 0.9 = 90 per cent. 
= the assumed efficiency of the heater. 

Note. — When the result obtained by For. (77) is a temperature 
greater than the temperature of the exhaust steam, it means that all of 
the steam will not be condensed. The temperature of the discharge 
from the heater is, then, within 2 to 5 deg. fahr. of the exhaust steam 
temperature, and the amount of steam condensed may be calculated by 
For. (78). 

Example. — A 1,200 h.p. condensing engine uses 20 lb. of steam per h.p. 
per hr. The auxiliaries use 2,400 lb. of steam per hr. The exhaust from 
the auxiliaries is condensed in a through-type open feed-water heater. 
The atmospheric relief -valve above the heater is set for a back-pressure of 
4 lb. per sq. in. The feed-water is delivered from the hot-well to the 
heater at a temperature of 110 deg. fahr. What is the temperature of 
the water flowing from the feed-water heater to the feed-pump? 

Solution. — The quantity of water delivered to the heater = (1,200 X 
20) = 24,000 lb. per hr. As given in a table of the properties of saturated 
steam, the total heat, above 32 deg. fahr., in steam at a pressure of 4 lb. 
per sq. in., gage, is 1,155 B.t.u. per lb. Hence, by For. (77), the temper- 
ature of the water leaving the heater = T f2 = [T/i W/ + 0.9W,(# + 32)]/ 
(W/ + 0.9W,) = {(110 X 24,000)+ [0.9 X 2,400 X (1,155+32)]} h- [2,4000 
+ (0.9 X 2,400)] = 199 deg. fahr. 

263. In A Non-Condensing Plant Only About One-Seventh 
Or Fourteen Per Cent. Of The Steam Exhausted From The 
Engine And Auxiliaries Can Be Utilized For Feed-Water 
Heating; About Eighty-Six Per Cent. Of The Exhaust Steam is 
Wasted. — The feed water should usually be heated to 212 deg. 
fahr. It is impossible to heat it to a higher temperature at 
atmospheric pressure without causing it to vaporize into steam. 
And, furthermore it is an impossibility to heat the feed water 



230 STEAM POWER PLANT AUXILIARIES [Div. 7 

to a temperature higher than that of the exhaust steam which 
is used for the heating. The temperature of this exhaust 
steam is always, at atmospheric pressure, 212 deg. fahr. 

Note. — The Exhaust From The Engine And Auxiliaries Is 
Practically All Steam, although it carries some condensed water. 
This exhaust steam holds the same amount of heat as any steam at 212 
deg. fahr. Now the latent heat in this steam, the heat which each pound 
of steam will give up in changing from steam at 212 deg. to water at 
212 deg. is, as taken from a steam table, 970.4 B.t.u. But the heat 
required to raise the temperature of 1 lb. of water from 50 deg. fahr. 
(which is the average cold feed-water temperature) to 212 deg. fahr. is 
only: 212 — 50 = 162 B.t.u. Therefore, the number of pounds of cold 
feed water which will be heated from 50 deg. fahr. to 212 deg. fahr. by 
1 lb. of exhaust steam will be 970.4 •*- 162 = 6 lb. One lb. of steam will, 
then, afford all of the heat that 6 lb. of feed water can, under the circum- 
stances, absorb. 

264. The Proportion Of The Total Steam Generated In A 
Non-Condensing Plant Which Is Useful In Feed Water Heat- 
ing is about 14 per cent. For each 6 lb. of cold water at 50 
deg. fahr. (as above described) which is pumped into the boiler 
1 lb. of water condensed from exhaust steam is pumped in with 
it. (This assumes that an open feed-water heater is used). 
This gives a total of 7 lb. of hot feed water pumped into the 
boiler for each pound of exhaust steam used. Thus (See also 
Sec. 263) only about pj or 14 per cent, of the total water pumped 
into the boiler (that is, H of the steam generated but finally 
exhausted through the engine and auxiliaries) can be effective 
for feed-water heating. The remainder, or 86 per cent, of the 
exhaust steam is wasted — unless it is employed for room heat- 
ing or some similar useful non-power-generation purpose. 

265. In A Condensing Plant Which Carries A 26-Inch 
Vacuum Only About One-Eleventh Or Nine Per Cent. Of 
The Steam Generated By The Boiler Can Be Used For Heat- 
ing The Feed Water. — In a condensing plant all of the steam 
from the engine is condensed with cold water and is discharged 
into the hot well. Some of the auxiliaries should be operated 
non-condensing so that their exhaust can be used for heating 
the feed water from the hot well up to 212 deg. fahr. if possible. 
The temperature of this condenser-discharge water which is 
thus used from the hot well for boiler feed is (with a 26-in. 



Sec. 266] FEED-WATER HEATERS 231 

vacuum) about 120 deg. fahr. Therefore, to raise its temper- 
ature to 212 deg. fahr. there will be required only 212-120 = 
92 B.t.u. It is assumed that an open feed- water heater 
will be used. Hence, for these conditions the number of 
pounds of feed water which will be heated from 120 deg. to 212 
deg. by 1 lb. of- exhaust steam (which will give up 970.4 B.t.u. 
of latent heat in changing from steam at 212 deg. to water at 
212 deg.) will be:— 970.4 + 92 = 10.6 lb. That is, 1 lb. of the 
exhaust steam at 212 deg. fahr. will heat 10.6 lb. of the 120 
deg. fahr. feed water to 212 deg. fahr. How, with each pound 
of the hot-well water which is fed into the boiler, the 1 lb. of 
condensed steam which is used in raising the temperature of 
the hot-well water is fed in with it. Hence, for each 1 lb. of 
exhaust steam utilized for feed-water heating there is fed into 
the boiler : — 10.6 + 1 = 11.6 lb. of feed water at a temperature 
of 212 deg. F. 

This being true, there is only 1/11.6 = 8.6 per cent, or say, 
9 per cent, of the total steam generated by the boiler which can 
be used for heating feed water. Obviously, then, the ideal 
economic condition for a condensing plant which carries a 26 
in. vacuum is to have auxiliaries which will furnish exhaust 
steam to an amount equivalent to about 9 per cent, of the 
steam generated by the boiler. It should be understood that 
the 9 per cent, is the ideal value which applies only for the 
water temperature conditions specified for this example. 
Losses such as condensation and the like, for which no allow- 
ance has been made in this problem, will tend to increase above 
9 per cent, the amount of exhaust steam which can be used for 
feed-water heating. 

Note. — It Is Reasonable To Expect That The Auxiliaries In 
The Average Plant Will Supply About The Amount Op Exhaust 
Steam Required For Heating The Feed Water. Every effort should 
be exerted to produce just enough exhaust steam to heat the feed water 
up to 210 deg. or 212 deg. But there should be no exhaust in excess of 
this. If there is excess exhaust the heat in it will be wasted. 

266. To Compute The Weight Of Steam Condensed By An 
Open Heater, use the following formula: 



232 



STEAM POWER PLANT AUXILIARIES 



[Div. 7 



Wherein: W s = weight of steam condensed, in pounds per 
hour. jP/2 = discharge temperature of feed- water in deg. fahr. 
Tfi = initial temperature of feed-water in deg. fahr. W F = 
weight of hot water delivered by heater in pounds per hour. 
H = the total heat in the exhaust steam, above 32 deg. 
fahr., in British thermal units per pound. 0.9 = 90 per cent, 
which is the assumed efficiency of the heater. 

Example. — Suppose 2,400 lb. per hr. of feed water is required by a 
boiler. Steam at 227 deg. fahr. is available for feed-water heating. 
The initial temperature of the feed-water is 90 deg. fahr. and it is delivered 
at 212 deg. fahr. What weight of steam is condensed by the heater? 

Solution. — As given in a table of the properties of saturated steam, 
the total heat above 32 deg. fahr., in steam at 227 deg. fahr. is 1,156 
B.t.u. per lb. By For. (78) the weight of steam condensed, 

(T f 2-T fl )W F 



W s = 



0.9(ff + 32) -T n +0.ir /2 

(212 - 90)2,400 



0.9(1,156 + 32) - 90 + (0.1 X 212) 



293 lb. per hr. 



,-Heact 



^•Crane Water Inlets- 
si- 



Live Steam. 
Connection'. 




Water Qutlet ' 
To Boilers--' 

Fig. 239. — Hoppes Live-Steam Heater And Purifier With Head Removed And 
Hanging On The Crane. (Heaters of very similar design are used for regular ex- 
haust-steam heating service.) 

267. The Pan Or Tray Area Required In An Open Heater 

using pans or trays (Fig. 239 and 240) is (Kent's Mechan- 
ical Engineers' Pocketbook) as follows: 



Quality of water 


Surface in sq. ft. per 1,000 lb. of water 
heated per hour 




For vertical type 


For horizontal type 


Very bad water 


8.5 
6.0 
2.0 


9 1 


Medium muddy water 

Clear water little scale 


6.5 
2 2 







Sec. 267] 



FEED-WATER HEATERS 



233 



Note. — The practice in heater manufacture is, however, to use a total 
tray surface equal to about 3 to 4 times the horizontal sectional area of 
the shell at the plane at which the trays are located. The space between 
the pans or trays is made not less than 0.1 the width for rectangular and 
0.25 times the diameter for round, trays or pans. It is not customary to 



Separating? Ba ffles •, 




Fig. 240. — Blake-Knowles Open Exhaust-Steam Feed-Water Heater Using Inclined 

Trays. 



use more than six pans in a tier. The size of the water storage or settling 
space in the horizontal type varies from 0.25 to 0.4 the volume of the 
shell; and in the vertical type from 0.4 to 0.6. The filters occupy from 
10 to 15 per cent, of the volume of the shell in the horizontal type and 
from 15 to 20 per cent, in the vertical. 



234 STEAM POWER PLANT AUXILIARIES [Div. 7 

268. To Compute The Approximate Size Of Shell Required 
For An Open Heater, use the following formulae: 

W f 

(79) A f = ^j~ (square feet) 

or 

(80) L h = J^ (feet) 

Wherein: A/ = cross-sectional area of heater, in square feet. 
L h = height of heater, in feet. W/ = weight of feed-water 
heated by the heater, in pounds per hour. K = a constant; 
for clear water K = 270; for slightly muddy water K = 200; 
and for very muddy water K = 70. The formula is based on 
proportions of commercial heaters furnishing 6,000 or more 
lbs. per hour of feed-water. These heaters were all of upright 
design having L h not more than 3 times the smaller base 
dimension. For heaters furnishing 3,000 to 5,000 lb. per hr., 
allow 25 per cent, more capacity than given by the formula. 

Example. — What should be the tray area and shell dimensions of an 
open heater to heat 10,000 lb. per hr. of feed water. The heater is to be 
square in cross section and the height is to be twice the base dimension. 
The water is slightly muddy. 

W f 10,000 

Solution. — By For. (80), L h = ——- = or A f L h = 50 cu. ft. 

A/K Ay X 200 

But, for a square section one side being ^L*, ]4,Lh X }4Lh X L* = 50 or 
L h = \/4: X 50 = 5.85 ft. or about 5 ft, 10 in. The base is 2 ft. 11 in. 
or 2.92 ft. square. The tray area required is (Sec. 267) about 3.5 times 
the cross-sectional area, or 2.92 X 2.92 X 3.5 = 30 sq. ft. approx. 
Assuming that six trays are to be used they will beabout -\/30/6 or 2 ft. 
3 in. square. They should be at least 0.1 the width or 2% in. apart. 



Sec. 2691 



FEED-WATER HEATERS 



235 



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236 



STEAM POWER PLANT AUXILIARIES 



[Div. 7 



270. Table Of General Data And Approximate Net Selling 
With Exhaust Steam (Harding and Willard, Mechanical 



Horsepower rating 



50 



100 



150 



200 



250 



300 



350 



425 



500 



Pounds feed water per hour. . . 

Weight in pounds 

Net price f.o.b. factory 

"Width, inches 

Depth, inches 

Height, inches 

Max. dia. exh. inlet and outlet 

Dia. cold water supply , 

Dia. ins. pump suction , 

Dia. waste and overflow 

Number of trays 

Length per tray, inches. ....... 

Width per tray, inches 



1500 

1200 

$102 

25 

21 

62 

4 

1 

IK 

1H 

4 
17 
12 



3000 

1300 

129 

27 

23 

63 

5 

1 

2 

1H 

4 

19 

13H 



4500 

1800 

159 

30 

25 

70 

6 

1H 

2H 

2 

4 

21 

15 



6000 

2100 

188 

32 

27 

73 

6 

IK 

2y 2 

2 

5 

22 

15 



7500 

2400 

229 

34 

29 

78 
7 

3 

2H 

5 

24 

16M 



9000 

2700 

256 

43 

29 

78 

7 

3 

2y 2 

5 

24 

16H 



10500 

3000 

275 

39 

33 

84 

8 

2 

4 

3 

5 

28 

21 



12750 

3300 

302 

49 

33 

84 

8 

2 

4 

3 

5 

28 

21 



15000 

3700 

331 

45 

38 

75 

9 

2 

4 

3 

10 

32 

10M 



Note. — The heaters tabulated above are designed for power-plant operation, and not 
See notes below Table 271 regarding prices. For estimating purposes and preliminary 
10 per cent to cover steam consumption of auxiliaries (pumps, etc.). The value so 
hour." Select a heater accordingly. In considering heaters of the same general type, 



271. Table Of General Data And Approximate Net Selling 
With Exhaust Steam. (Harding and Willard, Mechanical 



Horsepower rating 


50 


60 


70 80 


100 


1 
1301 160 


200 


240 


Pounds feed water per hour 


1500 


1800 


2100 


2400 


3000 


3900 


4800 


6000 


7200 




17 


20 


23 


?,7 


33 


43 


53 


67 


80 




18 


18 


18 
1V4 


18 


18 
IK 


36 

\K 


36 
IV* 


36 


36 


Diameter of tubes, inches 


Length of tubes* inches 


35% 


42V* 


49** 


56 V A 


70 


4534 


56 


695^ 


83 V4 




12 


12 
IV* 


12 


12 

m 


12 

m 


16 
2 


16 
2 


16 
2 


16 
2 


Diameter of feed pipe, inches 


Diameter of exhaust pipe, inches 


6 


6 


6 


6 


6 


8 


8 


8 


8 


Total length — horizontal heater 


4' 7" 


5' 2" 


5' 8" 


6' 3" 


7' 5" 


5' 7" 


6' 5" 


7' 7" 


8' 8" 


Total length f vertical type . . . 


5' 4" 


5' 11" 


6' 6" 


7'0" 


8' 2" 


6' 4" 


7' 2" 


8' 4" 


9' 5" 


on legs \ horizontal type 


2' 6" 


2' 6" 


2' 6" 


2' 6" 


2' 6" 


3'0" 


3'0" 


3'0" 


3'0" 


Shipping weight, / vertical type. . . 


880 


900 


950 


1000 


1125 


1250 


1550 


1700 


1900 


pounds 1 horizontal type 


950 


1000 


1050 


1250 


1400 


1675 


1750 


1900 


2000 


Net selling ( vertical type. . . 


$133 


140 


147 


154 


168 


193 


214 


235 


252 


price \ horizontal type 


$144 


155 


163 


171 


186 


214 


238 


260 


280 



Note. — "Closed" feed water heaters are either of the water-tube or steam-tube type, 
exhaust steam passing through the shell. In the latter the exhaust steam is passed 
shell. The water-tube heater is the type generally used in steam-power-plant work. 
Heaters may be vertical or horizontal type as space dictates. See note under Table 
note (a) material of tubes; (b) square feet of tube heating surface; (c) the weights; (d) the 

Note. — The prices listed above are for 1916 and cannot be relied upon closely at 



Sec. 271] 



FEED-WATER HEATERS 



237 



Prices Of Feed-Water Heaters Of The Open Type— For Use 

Equipment of Buildings, Vol. II) 



600 


750 


850 


1000 


1250 


1500 


1750 


2000 


2500 


3000 


4000 


5000 


6000 


18000 


22500 


25500 


30000 


37500 


45000 


52500 


60000 


75000 


90000 


120000 


150000 


180000 


4300 


4900 


5400 


6400 


7000 


8300 


9100 


10000 


11000 


12000 


15000 


16000 


17000 


380 


420 


493 


540 


618 


720 


820 


925 


1060 


1155 


1410 


1605 


1738 


55 


50 


60 


56 


68 


67 


78 


113 


113 


115 


128 


130 


132 


38 


42 


42 


48 


47 


56 


53 


42 


48 


54 


54 


62 


70 


75 


84 


84 


87 


84 


97 


97 


88 


88 


88 


100 


100 


100 


10 


10 


12 


12 


12 


14 


14 


16 


16 


18 


20 


22 


24 


2H 


2K 


2H 


2y 2 


3 


3 


3 


3K 


w* 


4 


4K 


5 


5K 


4 


4 


5 


5 


5 


5 


6 


6 


7 


8 


9 


10 


10 


3H 


3K 


3K 


BH 


4 


4 


4 


5 


5 


6 


7 


8 


8 


10 


10 


10 


20 


20 


20 


20 


20 


40 


40 


40 


40 


40 


32 


36 


36 


22 


22 


25 


25 


36 


22 


25 


25 


29 


33 


iok 


12 


12 


15 


15 


18 


18 


15 


15 


15 


18 


18 


18 



designed to operate in conjunction with steam-heating systems under back pressure, 
determinations, compute the steam consumption, per hour of the main engines, and add 
obtained corresponds to the line of the table entitled "pounds of feed water heated per 
but of different manufacture, compare particularly cubic contents, weights, and prices. 

Prices Of Feed-Water Heaters Of The Closed Type— For Use 

Equipment Of Buildings, Vol. II) 



300 


350 


400 


500 


600 


700 


800 


900 


1000 


1200 


1500 


1800 


2000 


9000 


10500 


12000 


15000 


18000 


21000 


24000 


27000 


30000 


36000 


45000 


54000 


60000 


100 


117 


133 


167 


200 


233 


266 


300 


333 


400 


500 


600 


667 


60 


60 


60 


90 


90 


90 


126 


126 


126 


126 


150 


150 


186 


IK 


IK 


IK 


IK 


IK 


IK 


IK 


IK 


IK 


IK 


1% 


1% 


IK 


62% 


73 


83K 


69K 


83% 


96% 


78K 


88% 


97% 


117% 


112% 


135 


111% 


21 


21 


21 


25 


25 


25 


29 


29 


29 


29 


34 


34 


39 


2H 


2K 


2y 2 


3 


3 


3 


4 


4 


4 


4 


5 


5 


6 


10 


10 


10 


12 


12 


12 


16 


16 


16 


16 


18 


18 


22 


V 3" 


8'1" 


9'0" 


8' 2" 


9' 4" 


10' 5" 


9' 2" 


10' 1" 


10' 10" 


12' 5" 


13' 5" 


15' 3" 


13' 10" 


8' 9" 


9' 7" 


10' 5" 


9' 7" 


10' 8" 


11' 10" 


10' 8" 


11' 7" 


12' 4" 


14' 0" 


14' 4" 


14' 2" 


13' 8" 


3' 5" 


3' 5" 


3' 5" 


3' 10" 


3' 10" 


3' 10" 


4' 3" 


4' 3" 


4' 3" 


4' 3" 


4' 11" 


4' 11" 


5' 6" 


2500 


2800 


2900 


3800 


4000 


4400 


5000 


5500 


5800 


6300 


7200 


11000 


14000 


2600 


2900 


3200 


4100 


4300 


4700 


5500 


6000 


6500 


7200 


10000 


12000 


14000 


322 


350 


378 


490 


540 


575 


660 


708 


750 


840 


1190 


1300 


1430 


356 


390 


420 


545 


598 


637 


730 


785 


830 


938 


1320 


1420 


1610 



In the former the feed water circulates through the tubes and is surrounded by the 
through the tubes and the feed water (surrounding the tubes) is carried through the 
The shell is usually of cast iron and brass or copper tubes are almost always used. 
270 regarding prices, and estimation and selection of heater. In making comparisons, 
prices, 
present. 



238 



STEAM POWER PLANT AUXILIARIES 



[Div. 7 



272. General Rules For Selecting Exhaust Steam Feed- 
Water Heaters are: Use an open heater whenever possible 
on account of its greater efficiency as a heater and purifier 
and ease of cleaning. It cannot be used: (1) When the 
feed-water in the heater must be under a pressure of more than 

about 5 lb. persq. in. (2) When 
the steam used for heating is ex- 
hausted under a vacuum as in 
condensing operation. (3) 
When the feed-water must be kept 
entirely free of oil. (4) When the 
feed-water heater is connected to 
the feed pump between the pump 
and the boilers. Under any of 
the four conditions listed a closed 
heater must be used. 



Note. — The Effectiveness Of An 
Open Feed-Water Heater As A 
Purifier depends not alone upon the 
area of heating surface which it con- 
tains, but also upon its volume of 
water-storage capacity. Storage ca- 
pacity is variable to a greater extent 
than is heating surface. If the water 
is hard, purification is desirable. The 
longer the water remains in the heater, 
the more thorough will be the pre- 
cipitation. Hence, a larger water- 
storage space is required than would 
otherwise be necessary. On the other 
in a surface-condensing plant, where 




Fig. 241.— The National Coil Type 
Closed Feed-Water Heater. 



hand, the heater may be used 
the condensate, which is usually free from scale-forming- impurities is 
used as feed water. Then, if there is a fairly uniform load, the con- 
densate is delivered to the heater at a uniform rate, and only such 
volume of water need be carried in storage as will insure a steady supply 
to the feed-pump. 

273. Closed Exhaust-Steam Feed-Water Heaters May Be 
Grouped Into Two Classes : (1) Water-tube heaters (Fig. 241) 
in which the feed-water passes through a set of brass or copper 
tubes which are surrounded by the exhaust steam. (2) 
Steam-tube heaters (Fig. 242) in which the exhaust steam 



Sec. 273] 



FEED-WATER HEATERS 



239 



HotJMxter 
Outlet-. 



Brass 
Steam 
Tubes 



Safety- 




^SWNWN ^ 



Cono/erisettion DripP/pe Connected 
To Steam Space 



Water Inlet- 

Steam Flow-. -Water Flow 




Fig. 243. — Diagram Of Parallel-Cur- 
rent Return-Flow Closed Feed-Water 
Heater. 



Water Outlet- 



Steam 




Water Inlet— -r^VSSteam'; 
Outlet- 



Fig. 242.— Steam-Tube Type Of Closed Fig. 244.— Diagram Of Counter-Cur- 

Exhaust-Steam Feed-Water Heater. rent Return-Flow Closed Feed-Water 

Heater. 




I-Section CC 
{Water Flow) 



I-Section 
(5team Flow) 



HT-5 action k A 
(Water Flow) 



Fig. 245. — Cross Sections Of Blake-Knowles Heater (Fig. 220) Showing Multi-Flow 

Arrangement. 



240 



STEAM POWER PLANT AUXILIARIES 



[Div. 7 



passes through a set of brass or copper tubes which are sur- 
rounded by the feed water. 

Note. — Closed feed-water heaters may be designed so that (Fig. 243) 
the water and steam flow in the same direction, or (Fig. 244) in opposite 
directions. In the first case, the heater is called a parallel-current heater. 
In the second case it is called a 
counter-current heater. If the 
heater is so built that the water 
flows straight through, it is called 
a single-flow heater. If the water 
flows back and forth through the 
tubes a number of times (Figs. 220 
and 245) it is called a multi-flow 
heater. If the water flows through 
coiled tubes (Fig. 241) it is called 
a coil heater. If the water is forced 
across the heating surface in a thin 
sheet or film it is called a film 
heater. 

274. The Tubes In A Closed 
Feed -Water Heater May Be 
Either Straight (Fig. 220) Or 
Spirally Corrugated (Fig. 
246). It is claimed for the 
corrugated construction that 
the spiral flow of the water 



-Sect/on Of Cast-iron Head 
Copper Corrugated Tube- 




ferrule For Expanding 
Tube-End 




Baffle 
Plate- 



Draining Pipe 



Floating Head--' 
Drain Opening 



Fig. 246. — End Of Copper Corru- 
gated Tube In Wainwright Closed Feed- 
Water Heater. 



Fig. 247. — Sehutte and Koerting Ver- 
tical Straight-Tube Closed Heater — Multi- 
Flow Type. 



through the tubes increases the contact pressure between 
the water and the tube surface, thereby facilitating the heat- 
transmission. It is also claimed that the spiral currents of 
water tend to scour the surfaces and prevent the accumulation 
of scale thereon. 



Sec. 275] 



FEED- \VA TER HE A TERS 



241 



275. A Corrugated Heater-Tube Gives Greater Heating 
Surface, for a given water volume, than does an uncorru- 
gated tube. Further advantages claimed for corrugated 
tubes (Fig. 246) are: they give a higher rate of conduction 
per unit length than smooth tubes, the corrugations take 
up all heat strains making more rigid construction of the 
heater possible. Corrugated tubes, it is claimed, are prefer- 
able where the range of temperature of the water, between 
inlet and outlet, is extreme, or where the velocity of the water 
through the heater is very high. 

Note. — When Straight Uncorrugated Tubes Are Used In Closed 
Heaters, a floating head arrangement (H, Fig. 247) is usually used to allow 



■■Double Coil Of Copper 
Tubing 




Fig. 



248. — Inside Manifold Of Whitlock 
Double-Coil Closed Heater. 




Wrought- 
Iron 
Strap 



Fig. 249.— How The Coils Are 
Secured In A National Coil 
Heater. 



for expansion in the tubes. Where the tubes are bent (Fig. 242) or 
coiled (Fig. 241) this feature is unnecessary as the tubes are free to 
expand and contract without straining the supporting head. Methods 
of connecting and supporting coiled tubes are shown in Figs. 248 and 249. 

276. Steam-Tube Closed Feed-Water Heaters (Fig. 242) 
are designed for service where a varying demand for steam 
necessitates very irregular feeding of the boilers. This condi- 
tion might exist where the steam-using apparatus which 
supply the exhaust steam for heating the water are operated 
intermittently. By sending the steam, instead of the water, 
through the tubes, the space surrounding the tubes is available 
for storage of a comparatively large volume of heated water 

16 



242 



STEAM POWER PLANT AUXILIARIES 



[Div. 7 



during the intervals when the feed-valve is closed. The 
water stored in the heater may then absorb the heat from 
the intermittent deliveries of exhaust steam. 

Note. — An Intermittent Delivery Of Exhaust Steam To The 
Feed- Water Heater might occur in a plant where hydraulic-elevator 
pumps, or the engines for operating trip-hammers, cotton compresses, 
or in similar irregular service, are depended upon for supplying the steam. 

277. To Compute The Tube Heating-Surface Required For 
A Closed Exhaust-Steam Feed-Water Heater, use the follow- 
ing formula: 

W,(!F /2 - T n ) 



(81) 



A f = 



u(T fs - 



T n + T 



fv 



(square feet) 



Wherein A f 

1000 



the total heating surface of the tubes, in square 
feet. W/ = the weight of feed 
water to be heated, in pounds 
per hour. T/i = the tempera- 
ture of the water entering the 
heater, in degrees Fahrenheit. 
Tf 2 = the temperature of the 
water leaving the heater, in de- 
grees Fahrenheit. T/ s = the 
temperature of the exhaust 
steam in degrees Fahrenheit. 
U = the coefficient of heat- 
transfer for the surface, in 
British thermal units per hour 
per square foot per degree tem- 

Fig. 250.-Graph Showing Effect Of Derafure difference as ffiven ill 

Water Velocity On Coefficient Of Heat P era ™ re Ctmerence as given 111 

Transfer Through Tubes Of Closed Table 278. 
Feed-Water Heaters. 

Note. — The Coefficient Of Heat Transfer In Closed Heaters 
Varies Within Wide Limits. It depends mainly upon the thickness 
and composition of the conducting wall, the disposition of the heating- 
surface, the water velocity through the heater (Fig. 250) and upon the 
conditions under which the heater is operated. It may range from 150 to 
1000 or more. The first of these values may be obtained with a steel- 
tube heater in which the water-velocity is low. The second may be 
realized with corrugated brass-tube heaters, of the film type, in which 




Water Yeloaty-Ft. Per Mm 



Sec. 278] 



FEED-WATER HEATERS 



243 



the water-velocity is very high. The values given in Table 278 are for 
commercial designs 

Example. — A closed exhaust-steam feed-water heater is required to 
heat 10,000 lb. of feed-water per hr. from 60 to 196 deg. fahr. with steam 
at 212 deg. fahr. The heater is to be of the multi-flow corrugated brass- 
tube type. What should be the area of the tubing? 

Solution:— By Table 278 U = 400. Hence, by For. (81) A f = W f 
(T f2 - T fl )/(U{T /s - V 2 [T fl + T f2 }}) = 10,000 X (196 - 60)/ 
(400(212 - ^[60 + 196]}) = 40.5 sq. ft. 

Note. — Increasing the velocity of the water passing through a heater 
increases (Fig. 250) the coefficient of heat transmission. In order to 
realize the possible maximum feed-water temperature, and at the same 
time use a moderately high velocity of flow, the tubes should be as 
long as is feasible, and of small diameter. 

278. Table Showing Average Coefficients Of Heat Trans- 
mission In Closed Feed-Water Heaters (Gebhardt). 



Type of heater 


Average coefficient 

of heat-transfer = 

U, For. (81) 


Single-flow, steel water-tube 


150 


Single-flow, plain brass water-tube . . 


200 


Single-flow, corrugated brass water-tube 


300 


Spiral coil, plain brass water-tube 


350 to 700 


Multi-flow, plain brass water-tube 


350 


Multi-flow, corrugated brass water-tube 


400 


Multi-flow, plain brass water-tube, with retarders 
Film, corrugated water-tubes 


450 
600 


Multi-flow, iron steam-tube 


100 to 225 


Multi-flow, brass steam-tube 


200 to 450 


Multi-flow copper steam-tube 


220 to 475 







279. Closed Exhaust-Steam Feed-Water Heaters Are 
Sometimes Rated In Terms Of Heater Horsepower. — By 

using For. (81) it can be shown that one square foot of heater 
surface will suffice to heat 103.5 lb. of water per hr. from 60 
deg. fahr. to 194 deg. fahr., with a coefficient of heat transfer 
(Sec. 277) of about 165 B.t.u. per sq. ft. per hour per deg. 
difference in temperature. On the above outlined basis and 
on the assumption that 34.5 lb. of feed water is required per 
boiler horse-power per hour, a closed heater will supply 
103.5 -5- 34.5 = 3 boiler horse power per sq. ft. of heating 



244 



STEAM POWER PLANT AUXILIARIES 



[Div. 7 



surface. Hence: 1-f 3 = ^ sq. ft. of heater surface is 
sometimes allowed per boiler horsepower. 

Note. — Double Heater Installations (Fig. 251) are used in large 
power plants which are operated continuously. These consist of two 
separate feed-water heaters which are so connected as to receive exhaust 
steam from a common exhaust pipe and water from a common water- 
supply pipe. With an installation of this kind, one heater may be cut 



Exhaust Outlet To Atmosphere—-.^ 
Oil Separator- ^ 




-Colo/ Water Regulating Va/ve 
I - P I a n 

Exhaust Outlet To Atmosphere--. 



Water Inlet- 



f>0| 



Water Inlet 




^#%^^ 



Overflow 



Veed-Pump Suction 

Fig. 251. — A Double Heater Installation 



Main Exhaust Inlet-' 
I-Elcvcttioh 



out of service for cleaning or inspection while the other continues to 
supply hot water for the boilers. 

280. The Relative Advantages And Disadvantages Of 
Open And Closed Feed-Water Heater may (See Gebhardt's 
Steam Power Plant Engineering) be summed up as follows: 
(1) With an open heater the water may be heated to the temperature 
of the exhaust steam. With a closed heater the possible maximum 



Sec 281.] FEED-WATER HEATERS 245 

temperature of the feed-water will always be less than the temper- 
ature of the exhaust steam. (2) Ordinarily, the pressure in 
an open heater is but slightly in excess of atmospheric pressure. 
Ordinarily, the pressure of the water in a closed heater is some- 
what in excess of boiler pressure. (3) An open heater is liable 
to rupture by the building up of a back-pressure, due to sticking 
of the back-pressure valves. A closed heater is built to with- 
stand any pressure which is likely to occur. (4) Oil in the exhaust 
steam may contaminate the feed-water in an open heater. Oil 
cannot enter the feed-water from the exhaust steam in a closed 
heater. (5) Scale and other impurities precipitated in an open 
heater are readily removed. It is difficult to remove scale from a 
closed heater. If the feed water contains a high content of 
scale-forming impurities, then, usually, the open heater is the 
preferable and in some eases the only permissible type. (6) 
An open heater must be located above the pump suction. The 
feed pump must be between the heater and the boiler. A closed 
heater may be located anywhere between the feed pump and the 
boilers. (7) Where the water is taken from a natural source of 
supply, two pumps are necessary with an open heater. With a 
closed heater only one pump is required in any case. (8) With 
an open heater the feed pump handles hot water. With a closed 
heater the feed pump handles cool water. (9) An open heater 
cannot be installed in the exhaust line from a condensing engine 
as can a closed heater. (10) The returns from a heating system 
cannot be delivered directly to a closed heater as to an open 
heater. 

281. In The Installation Of An Open Feed-Water Heater 
the following general directions should be observed. 

Directions. — (1) Locate the heater so that it may be conveniently 
piped to the source at the exhaust-steam supply and will be, at the same 
time, as close as possible to the boilers. 

(2) Set the heater plumb on a substantial foundation (Fig. 230) of 
proper height to bring the hot-water outlet to the feed-pump at least 
4 ft. above the discharge-valve deck of the pump. 

(3) Locate the feed-pump (Fig. 228) as close as possible to the heater. 
Also, run the suction pipe, of a size equal to the outlet orifice of the 
heater, as directly as possible. If the pump must be located at some 
distance from the heater, or the suction connection must be made with 
a number of sharp turns, either the suction pipe should be of larger 



246 STEAM POWER PLANT AUXILIARIES [Div. 7 

size than the outlet orifice of the heater or the heater should be set at a 
greater height than 4 ft. above the discharge-valve deck of the pump. 
In some cases both of these alternatives may be desirable. 

(4) Before connecting the mechanism of the float (H, Fig. 235) to the 
controlling valve in the water-supply pipe, see that the float and mechan- 
ism move freely. 

(5) Pack the filtering material, excelsior and coke (Fig. 235) closely 
between the filter plates. 

(6) If the heater is to be connected up for thoroughfare service (Fig. 
224) attach the engine exhaust pipe directly to the heater exhaust inlet, 
and, from the heater exhaust outlet, run a pipe to the atmosphere. This 
pipe should be of the same size as the exhaust outlet. A back-pressure 
or exhaust-relief valve should be placed in it at a point somewhere beyond 
any branch connection which may be made for supplying a heating 
system, or for other purposes. 

(7) See that the back-pressure valve (V, Fig. 229) is set for a pressure 
not higher than that which the heater will safely sustain. 

(8) If the heater is to be connected up for induction service (Fig. 228) 
run a branch from the main exhaust pipe to the heater exhaust inlet. 
This branch should be of the same size as the inlet orifice of the heater. 
It should contain a gate valve, so that the heater may be cut out for 
cleaning, and also to provide a means for regulating the supply of exhaust 
steam delivered to the heater. 

(9) A vent pipe (V, Fig. 230) should be attached to the top of an induc- 
tion heater. This is to allow air to escape and to insure admission of the 
requisite quantity of steam. The vent pipe may be screwed into a reducer 
flange bolted to the heater exhaust outlet. It should have a free opening 
throughout its length. The valve, V, (Fig. 230) in the vent pipe should 
never be closed except when the heater is cut out of service for cleaning. 

(10) Place a gate valve in the cold water supply pipe (F, Fig. 235) just 
beyond the controlling valve. Also place a gate valve in the pump suc- 
tion pipe. 

(11) Connect the oil-drip (W, Fig. 235) and the blow-off pipe to the 
sewer independently of each other. (If it is desired to recover and filter 
the oil for further use, the oil-drip may be piped to a separate reservoir.) 

(12) Cover the heater with asbestos, magnesia, or other heat-insulating 
substance, to prevent radiation of heat therefrom. 

282. In The Operation Of An Open Feed -Water Heater 

the following general directions should be observed. 

Directions. — (1) When the heater is first put in service, the cold-water 
controlling valve should be blocked open, also the blow-off valve should 
be opened, and a current of water permitted to run through until the 
heater is thoroughly flushed out. The blocking may then be removed 
from the controlling valve and the blow-off valve closed. 

(2) The lengths of the float connections should be so adjusted that the 



Sec. 283] FEED-WATER HEATERS 247 

controlling valve will, respectively, be open fully and closed tightly at the 
predetermined low and high water levels. 

(3) The blow-off valve should be opened once a day to blow out the 
sediment which may have collected in the bottom of the heater. 

(4) About once a week, or oftener, if necessary, the coke filter bed 
should be flushed out by opening the blow-off valve and admitting water 
under pressure through the flushing pipe (B, Fig. 235). 

(5) The pans of an open heater of the type shown in Fig. 239 should be 
removed and cleaned whenever the depth of scale is sufficient to interfere 
with their operation as settling basins. The time allowable before this is 
necessary depends on the nature of the water. 

283. In The Installation and Operation Of A Closed Feed- 
Water Heater the following general directions should be 
observed : 

Directions. — (1) The heater should be connected to the main ex- 
haust pipe as near the engine as may be practicable. 

(2) All feed-water and blow-off connections should be made with 
either box or flange unions, so that the parts can be easily taken apart for 
inspection. 

(3) A straightway valve or plug cock should be inserted in the blow-off 
pipe. 

(4) The safety valve on the feed-pipe, in the case of a water-tube 
heater, or on the heater itself, in the case of a steam-tube heater, should 
be loaded from 15 to 20 lb. per sq. in. above the boiler pressure. No 
other valve, of any kind, should be placed between the safety valve and 
heater. 

(5) The drip pipes should be of the same size as the drain orifices in 
the heater. The drip pipes should contain as few bends as possible and 
should incline downwards from the heater in all parts of their lengths. 

(6) The heater should be covered with a heat insulating material to 
prevent loss of heat by radiation. 

(7) The blow-off valve should be opened once a day to relieve the 
heater of any sediment that may have collected. 

(8) When the plant is shut down in cold weather, the heater should be 
thoroughly drained, to obviate danger of freezing. 

284. If The Safety Valve of a Closed Feed-Water Heater 
Will Not Remain Tight under the normal operating pressure 
it should be examined carefully to determine the cause. If 
the valve is of the lever type, extra weights should not be 
added to it in an effort to make it tight. Disaster may result 
from such procedure. Neither should the tension of the 
spring be increased, if it is of the spring type, unless an in- 
vestigation shows that the spring-tension is too low. If the 



248 



STEAM POWER PLANT AUXILIARIES 



[Div. 7 



valve does not close tightly after blowing off or if it " simmers " 
instead of blowing, it usually means that the seat or valve 
is in bad condition or that the adjusting ring is so far from 
the proper position that the valve is "out of control. " 

285. To Get The Most Effective Service From A Feed- 
Water Heater, It Must Be Cleaned at regular and frequent 
intervals. Local conditions must, in every case, determine 
the frequency of the cleanings. But in no case should the 
heater be operated longer than a month without cleaning. 

286. Live-Steam Heaters And Purifiers (Fig. 239) are 
intended, primarily, to purify the feed-water. They are 



Live Steam Purifier-. 

Exhaust 
Atmosphere \ 

Discharge 

Purifier- 



Direct Steam-Supply To Pump 

■Steam Supply To Pump Through Purifier 

Gravity feed From Purifier To Boilers 




"Open Exhaust- 
Steam Heater 



To Pump 



'Peed Pump "By-Pass For Direct Feed lb Boilers 



Fig. 252. — Hoppes Live-Steam Purifier Installed In Connection With Exhaust-Steam 
Heater. (When the purifier is in operation, the pump is supplied with steam 
through connection F in order that air and non-condensable gases liberated from the 
feed- water may be removed from the purifier.) 

installed (Fig. 252) where the feed-water contains scale- 
forming elements, as the sulphates of lime and magnesia, 
which precipitate at much higher temperatures than are 
obtainable in open exhaust-steam heaters. 

Note. — All Op The Scale-Forming Impurities Dissolved In A 
Feed- Water May Usually Be Precipitated In A Lwe-Steam Puri- 
fier if the water is properly pre-heated; see the author's Steam Boilers. 
The sulphates of lime and magnesia precipitate at temperatures above 
250 deg. fahr. The carbonates precipitate at a much lower temperature. 
It is claimed that 80 per cent, of the sulphates will, at a temperature of 



Sec. 2S6] FEED-WATER HEATERS 249 

250 deg. fahr., be deposited in a live-steam purifier. Also, that at a 
temperature of 300 deg. fahr., all of the sulphates will be deposited in the 
purifier. 

QUESTIONS ON DIVISION 7 

1. What are the three principal reasons why a feed-water heater should be used? 

2. Give an approximate rule for estimating the saving due to heating feed-water. 
Give an approximate rule for estimating the increase in boiler capacity due to feed- 
water heating. 

3. How does a feed-water heater protect a boiler from undue strains in the seams? 
Give an estimated value for heat strains in boiler plates caused by cold water in a 
boiler. 

4. What is an open heater? A closed heater? An atmospheric heater? A vacuum 
heater? 

5. What kind of heaters are used as primary heaters? How are they connected to 
other equipment? What are the approximate temperatures in a primary heater with 
good condenser action? What condition of the condenser and cooling water makes the 
use of a primary heater advisable? 

6. What is a secondary heater? How is it connected to the primary heater. What 
are the average temperatures for an open atmospheric heater steam supply and water 
outlet? 

7. What is an induction heater? A through heater? How is each piped? 

8. What operating condition governs the temperature of the exhaust steam available 
for use in a heater? 

9. Why is the feed-water temperature obtained with a closed heater ordinarily lower 
than that obtained with an open heater? 

10. What functions are performed by an ordinary exhaust-steam feed- water heater? 
Describe the operation of an open heater in detail. 

11. Name several scale-forming impurities that may be precipitated in an open 
heater. 

12. Why is all oil objectionable in feed-water? Why is cheap oil likely to be especially 
objectionable? 

13. What common dissolved gases are objectionable in feed water? Why? 

14. What is a water-tube heater? A steam-tube heater? A parallel-current heater? 
A counter-current heater? A single-flow heater? A multiflow heater? 

15. What is a coil heater? A film heater? 

16. What advantages are claimed for spirally corrugated heater-tubes? 

17. For what classes of service are steam-tube heaters particularly adapted? Why? 

18. What is the basis of the heater horsepower? 

19. What is a double-heater installation? 

20. What are the relative advantages and disadvantages of open and closed heaters? 

21. What is a live-steam purifier? 

22. What per cent, of the exhaust steam from a non-condensing engine does a feed- 
water heater ordinarily consume in heating the feed for the engine boilers. 

23. How does the volume of an open feed- water heater affect its efficiency as a purifier? 

24. Give a few general directions for the installation of an open feed-water heater. 

25. Give a few directions for preparing an open heater for service and keeping it 
working properly. 

26. Why should an oil separator usually be installed in the steam line to an open 
heater? 

PROBLEMS ON DIVISION 7 

1. Water at a temperature of 90 deg. fahr. is available for feeding the boilers in a 
power plant. The main engine runs condensing. It develops 500 h.p. on a steam- 
consumption of 20 lb. per h.p. per hr. The steam consumption of the auxiliaries is 
about 11 per cent, of that of the main engine. If the exhaust from the auxiliaries is 
condensed in an open atmospheric heater, what will be the temperature of the feed- 
water as delivered to the boilers? 



250 STEAM POWER PLANT AUXILIARIES [Div. 7 

2. A boiler generates steam at a pressure of 150 lb. per sq. in., gage. The water which 
is fed to the boiler is preheated with exhaust steam from 60 deg. fahr. to 210 deg. fahr. 
What saving of fuel results from thus utilizing the exhaust steam? 

3. The coal consumption of a set of boilers is 5 tons per day. The feed-water is 
delivered at a temperature of 150 deg. fahr. It is estimated that by using a quantity 
of exhaust steam which is now going to waste, the feed-water may be delivered at a tem- 
perature of 212 deg. fahr. The average steam pressure is 125 lb. per sq. in., gage. The 
fuel costs $3.50 per ton. The plant operates 300 days per year. It will cost about 
$300 to improve the present heating equipment. The rate of interest on the invest- 
ment is 6 per cent, per annum. The assumed rate of depreciation is 6.0 per cent, per 
annum. It will probably cost $4 per month to maintain and operate the apparatus. 
What will be the probable annual net saving? 

4. A closed exhaust-steam feed-water heater is required to heat 15,000 lb. of feed 
water per hr. from 70 to 200 deg. fahr. with steam at 220 deg. fahr. The heater is to 
be of the multiflow plain brass water-tube type. What should be the area of the tubing? 

6. If an open heater heats 15,000 lb. per hr. of feed- water from 40 deg. fahr. to 205 
deg. fahr. with steam at 212 deg. fahr., what weight of steam does it condense? 



DIVISION 8 

FUEL ECONOMIZERS 

287. A Fuel Economizer (Fig. 253) is an apparatus in which 
boiler feed-water is preheated by the combustion gases (Table 
288) which are discharged from boiler-settings. The econo- 
mizer is interposed in the path of the gases between the boiler 
and the chimney. 



.Stack 



Boilers - 




conomizer 



Fig. 253. — An Economizer Functions To Raise The Temperature Of The Feed Water. 
(Sturtevant Economizer Co.) 

288. Table Showing The Percentage Of The Heat Of 
The Fuel Which Is Present In The Gases Of Combustion As 
They Leave The Boiler {Green Economizer Co.). — Column A 
is based on an air supply of 18 lb., per pound of combustible. 
This represents average underfeed stoker operation with 
forced draft. Column B is based on an air supply of 24 lb., 
per pound of combustible. This represents average overfeed 
or natural-draft stoker operation. Column C is based on an 
air supply of 30 lb., per pound of combustible. This represents 
average operation with hand firing and natural draft. 

251 



252 



STEAM POWER PLANT AUXILIARIES 



[Div. 8 



Flue-gas temperature 


Per cent. 


3f heat of fuel in 


flue gases 


in deg. fahr. 


A 


B 


C 


300 






12.4 


350 




12.0 


14.9 


400 




14.0 


17.4 


450 


12.2 


16.1 


20.0 


500 


13.8 


18.2 


22.6 


550 


15.4 


20.3 


25.2 


600 


17.0 


22.4 


27.8 


650 


18.5 


24.4 


30.4 


700 


20.1 


26.5 




750 


21.7 






800 


23.2 







Note. — The Heat Which Is Utilized In An Economizer Does 
Not Represent A Clear Gain (Fig. 254). To compensate for the loss 
of natural draft, which results from lowering the chimney temperature, 
it is generally necessary to install a system of artificial draft. This 




Fig. 254. — Chart Showing Losses In Power Plant Operation. 



entails an extra expense for draft equipment, and for the subsequent 
operation and maintenance thereof. However, it is often profitable 
to install an economizer in a plant of greater capacity than about 500 
boiler horse power. 

289. There Are Two General Types Of Fuel Economizers : 

(1) The independent type (Figs. 255 and 256). (2) The integral 
type (Figs. 257 and 258). The first is located apart from the 



Sec. 289] 



FUEL ECONOMIZERS 



253 



Feed-Wafer Reversing $ears-^-- r ..^ Worms sDrivhg Shaft for Scrapers 



Outlet 




Fig. 255. — An Independent Fuel-Economizer. (Green Economizer Co.) 



■[noluced Draft 
Fan 




V///// ////////////A- /////////// //////// ///7///SS/ ////// //////A 




n-EI e v a t i o-'n 

Fig. 256. — A Typical Installation Of Independent Fuel-Economizers. (Hampton 
Mills, East Hampton, Mass.) 



254 



STEAM POWER PLANT AUXILIARIES [Dnr. 8 



boiler setting. The second is located within the boiler setting. 
Thus, it practically forms an integral part of the boiler 
structure. 




1 ^BEgg 



Fig. 257.— High- And Low-Pressure Economizers. (Kansas City Light And Power Co. ) 



290. An Independent Economizer (Fig. 255) consists, 
essentially, of a double series of cast-iron headers, or mani- 
folds (Fig. 259) which are connected together by vertical 
tubes. The tubes are commonly made of cast-iron. Their 
usual dimensions are 4% 6 -in, diameter and 9- to 12-ft. length. 



Sec. 291] 



FUEL ECONOMIZERS 



255 



The water, which is discharged by the boiler feed-pump, 
passes through the headers and tubes of the economizer before 
it enters the boiler. The hot gases, which flow from the 
boiler-setting to the chimney, pass (Figs. 260 and 261) through 
the spaces between the economizer tubes. The heat in the 
gases is thereby transmitted to the feed-water. 



,.-Prehevteror 
: Economizer 




Fig. 258. — Badenhausen Boiler Directly Connected With Integral Economizer Or 

Preheater. 



Note. — Economizer-Tubes May Be Arranged In Either Straight 
Or Staggered Rows. The staggered arrangement (Fig. 261) affords 
the greater facility for heat-transfer from the gases. The straight 
arrangement (Fig. 260) offers the least obstruction to the draft. Thus 
the advantage of either arrangement is apparently offset by the dis- 
advantage of the other. 

291. Integral Economizers are designed to withstand either 
high pressures or low pressures. High-pressure integral econo- 
mizers are so located (Figs. 257 and 258) as to receive the 
gases directly as they issue from contact with the boiler sur- 



256 



STEAM POWER PLANT AUXILIARIES 



[Drv. 8 



Draw-Out 
Rods-. 



; Top Header- 



loosened 
Tube 




Fig. 259. — Construction Of Headers And Tubes Of Sturtevant Economizer And 
Method Of Removal And Replacing Of Tubes. 



.-Tubes 



Tubes-. 




Direction of Gas Flow-- 



Fia. 260. — Economizer Tubes In 
Straight Rows. 




Direction of Oas Flow- 1 



Fig. 261. — Economizer Tubes In 
Staggered Rows. 



Sec. 202J FUEL ECONOMIZERS 257 

faces. These economizers are, therefore, built with wrought 
iron or steel tubes and drums. Low-pressure integral econo- 
mizers are so located (Fig. 257) as to receive the gases at a 
comparatively low temperature. Hence, these economizers 
are usually built, similarly to the independent type of econo- 
mizer (Sec. 289), with cast-iron tubes and headers. 

292. Certain Advantages And Disadvantages Attend The 
Use Of Cast-iron, Wrought Iron And Steel In Economizer 
Construction (Sees. 290 and 291). — Cast-iron tubes and headers 
are less susceptible to corrosion than are those which are made 
of wrought iron or steel. But the liability of cast-iron tubes 
and headers to fail under the stresses of expansion and con- 
traction, and pressure is by far the greater. 

Note. — Corrosion Of Economizers may be due, internally, to an 
acid property of the feed-water. Externally it may be due to sulphurous 
acid or dilute sulphuric acid which are formed by the action of moisture 
and SO2 in the sooty deposits on the tubes. The moisture may come 
from leaky joints, or it may be due to a sweated condition of the tubes. 

Sweating Of Economizer-Tubes occurs when the temperature of the 
metal falls below the dew-point of the combustion gases. This condition 
will generally result when water at a temperature less than about 130 deg. 
fahr. is pumped through the economizer. Certain economizer manufac- 
turers recommend that the entering feed-water temperature should 
be at least 90 to 100 deg. fahr. If the available feed water is colder, a 
by-pass may be arranged to pass some hot water into the feed line. 

293. Cleanliness Of The Tube -Surfaces, Both Inside and 
Outside, Is Essential To The Effectiveness Of A Fuel Econo- 
mizer. — The soot which is mingled with combustion-gases 
adheres very readily to economizer-tubes. This is due to the 
comparatively low temperature of the tubes. Soot is an excep- 
tionally poor heat-conductor. Hence the urgent necessity for 
its removal from the tubes is apparent. 

294. Two Methods Are Available For Removing Soot From 
Economizer-Tubes: (1) Scraping. (2) Blowing. The scrap- 
ing-method (Figs. 255 and 262) is the more commonly used. 

Explanation. — Economizer-tube scrapers (Fig. 263) are in the form of 
sleeves which encircle the tubes. These sleeves are caused to traverse 
the tubes, from end to end, by means (Figs. 255 and 262) of a geared 
mechanism. The soot, which is scraped off by the beveled edges of the 
sleeves, falls into a pit beneath the economizer. It is then removed 

17 



258 



STEAM POWER PLANT AUXILIARIES 



[Div. 8 



through clean-out doors. Or, the soot may drop into a pit (Fig. 255) 
whence it is conveyed away through a pipe. 

Notes. — Economizer Soot-Blowers (Fig. 264) are of the same type 
as those which are used with water-tube boilers. These blowers are 



Reversing Lever- 
Operating 
Pawls -,-... 



.'■Driving 
[ Pulley 




Fig. 262. — Mechanism For Automatic Reversal 
Of Travel Of Soot-Scrapers On "Green" Fuel- 
Economizers. Bevel-Gear Pinions B\ And Bi 
Are Loose On Shaft S. 




Fig. 263. — Soot-Scraper For 
E conomizer-Tubes. 



described in the author's Steam Boilers. It is claimed that they 
remove the soot entirely from the tube-surfaces. With the use of 
sleeve-scrapers (Fig. 263) a thin, compact, film of soot may constantly 
remain on the surfaces. 



..•Flow of Gases 



Sprocket * . , Blower Elements-, 




Soot 
'■Cleaner *•■ , „ , 

Header $ Drain Valve-- 

■■'.•'■' I - $ i d e V i e w ' • '.! ; ; • ' ' • %%//. 



11 



P.- End 'View 



Fig. 264. — "Green" Fuel-Economizer Equipped With Vulcan Soot Blowers. 

The Power Expended In The Operation Of Economizer-Tube 
Scrapers may be approximately 1 h.p. per 1000 sq. ft. of economizer 
surface. 

The Steam-Consumption Of A Soot-Blowing System depends upon 
the size of the system and the time-interval during which it must be used 



Sec. 2951 



FUEL ECONOMIZERS 



259 



to effectually remove the soot. A system consisting of six blower-units, 
each fitted with 38 nozzles, will consume 2600 lb. of steam during a blow- 
ing period of six minutes (Power House; July 5, 1919, p. 272). 

295. Deposits Of Scale And Sediment In Economizer-Tubes 
Are Detrimental To Economy (Fig. 265). — Where the feed- 
water contains scale- and mud-forming impurities, the econo- 
mizer should be frequently blown down. Also, the tubes 
should be washed out, as often as is necessary, with a hose. 
Formation of hard scale may, by these means, be prevented. 



240 
a: 
































































IE 

°" 2.26 










\ 




















A 


































a!_ 
























/ 




s 


J 


"> 


\ 














^s 








k 




XI 

E 3 2,20 

c o 

o =c 


























7 






























A 


































































^ 


§" w 715 


_E 


l' 




























































E a l]Q 


N 












\ ti 












































B" 


vt 




























\ 
















































° 100 













I 2 3 4 5 6 1 8 9 10 II 12 13 14 15 16 IT 18 19 20 21 2223 24 25 26 27 28 29 30 31 
Days of Month 

Fig. 265. — Diagram Showing Daily Average Consumption Of Coal When Econ- 
omizer Tubes Were Clean And When They Were Lined With Scale. Graph A-A 
Shows Consumption When Tubes Are Lined With Scale. A'- A' Is The Average Of 
A-A. Graph B-B Results When Tubes Are Clean. B'-B' Is The Average of B-B. 

Note. — Scale Does Not Form As Readily In Economizers As 
In Boilers. This is due to the lower temperature of the water in 
economizers. The temperature is, however, usually high enough to 
cause precipitation of sedimental impurities. The comparatively-low 
velocity of the water-flow in an economizer facilitates such precipitation. 
Hence the sediment readily settles into the bottom headers, whence it 
may be blown out through the blow-off valves. 

296. An Economizer Should Be Fitted With Instruments 
For Showing The Combustion-Gas And Feed-Water Tem- 
peratures (Table 301). — Thermometers should be inserted in 
the feed-water connections to the economizer, both at inlet 
and outlet. Also, a pyrometer should be inserted, at each 
end of the economizer, in the path of the combustion-gases. 
These instruments afford a ready means for checking the 
performance of the economizer. 

Explanation. — Suppose the instruments were to show a steady 
increase, above normal, of the flue-gas temperature at exit from the 



260 



STEAM POWER PLANT AUXILIARIES 



[Dtv. 8 



economizer, while the temperature of the outgoing feed water steadily 
decreases. This condition would probably indicate that the heat is 
excluded, by a steadily increasing coating of soot, from the tube surfaces. 

297. Infiltration Of Air Through The Setting Of An Econo- 
mizer Is Detrimental To Economy. — Cool air, passing in through 
crevices in the setting, will mingle with the current of combus- 
tion gases. The air will thereby absorb heat from the gases. 
Hence the quantity of heat delivered to the water, flowing 
through the economizer, will be diminished. 

Note. — Leakage of air into an economizer setting may occur where the 
tubes are cleaned with scrapers. The openings through which the 
scraper-chains pass may afford ready ingress for air. This difficulty does 
not attend the use of blowers. 

298. Excessive Leakage Of Air Into An Economizer Setting 
May Be Detected By Observing The C0 2 Drop Through The 
Economizer. — A drop of about 2 per cent, may reasonably be 



18 

17 

w 16 

o 

14 

« !J 

ti l2 






































































































































i 




















































I 
























5 10 

C 9 

— 8 

V* 

% 3 
u 2 




I 


























\ 


























\ 














































































































































































































a. ' 





























10 20 30 "40 50 60 10 80 90 100 110 120 130 
Lb. of Waste Gas per Lb. of Cooii 

Fig. 266. — Chart Showing The Pounds Of Gas Per Pound Of Illinois Coal Corresponding 
To Percentages Of CO2. (Power Plant Engineering, Apr. 1, 1919.) 

expected. When this percentage of drop is exceeded, the 
leakage of air is probably excessive. 

Example. — The combustion-gases, passing from a boiler, have a 
temperature of 600 deg. fahr. and contain 10 per cent, of CO2 as they 
enter an economizer. As they leave the economizer, due to infiltration 
of air, the gases have a temperature of 300 deg. fahr. and contain 6 per 
cent, of C0 2 . The outside-air temperature is 70 deg. fahr. The specific 



Sec. 298] 



FUEL ECONOMIZERS 



261 



^^^^^^^^^^f? 



OP; 
P. 




262 



STEAM POWER PLANT AUXILIARIES 



[Div. 8 



heat of the gases = 0.24. It is assumed (Fig. 266) that 1 lb. of coal 
yields 15.5 lb. of combustion gases when the CO2 amounts to 10 per cent., 
and 26 lb. of gases when the CO2 amounts to 6 per cent. What is the 
percentage of heat-loss? 

Solution. — The heat, above 70 deg. fahr., which is contained in the gases 
as they enter the economizer = (600 - 70) X 15.5 X 0.24 = 1,971.6 
B.t.u. per lb. of coal burned on the grate. The infiltration of air amounts to 
26 — 15.5 = 10.5 lb. per lb. of coal burned. The heat required to raise the 
temperature of the infiltered air to 300 deg. fahr. = (300 — 70) X 10.5 X 
0.24 = 579.6 B.t.u. per lb. of coal burned. Hence, the percentage of 
heat-loss = (579 -r- 1971.6) X 100 = 29.4 per cent. 

299. The Draft-Pressure Drop Through An Economizer 

(Fig. 267) depends upon the arrangement (Sec. 290) of the 
economizer-tubes, the velocity of the gases, and, perhaps, 



Economizer 



Draff, Inches of Wafer 
3J5 0.20 0.25 0.30 0,35 





















Vptake 


































Economiier 


















Entrance 


















1st. Poiss 


































2nd Pass 


































Fv'it 




















''■Induced (Superheater rDrum 
Draft 
Fan 



Boiler 
Tubes- 




m^^^^^^z^^^ 



Fig. 268. — Draft Pressure Drop Fig. 269. — Diagram Showing Arrange- 

Through 8,500 Sq. Ft., 3 Section ment Of Forced Draft And Induced Draft 

Economizer-Fan Draft. Fans In Connection With Economizer. 

(B. F. Sturtevant Co.) 

upon conditions peculiar to each installation. It may (Fig. 
268) vary from 0.15 to more than 0.3 in. of water column. 
The frictional resistance of the tubes is directly proportional 
to the length of the economizer and to the square of the 
velocity of the gases. 

Note. — Economizers Generally Prove Unprofitable Where 
Chimneys Are Alone Depended Upon To Create Draft Pressure. 
Cooling of the flue gases (Table 300) by the economizer diminishes the 
draft-producing effectiveness of the gases. In addition to the reduction 
of natural draft, due to this cause, there is the loss of draft-pressure due 
to pushing the gases against the frictional resistance of the economizer. 



Sec. 300] 



FUEL ECONOMIZERS 



263 



The cost of the additional chimney-height, necessary to compensate for 
these deficiencies, will often more than offset the possible gain due to 
heating the feed-water. 

Artificial Draft Is Generally Used With Economizer Instal- 
lations. The draft may be either forced (Fig. 269) or induced. Sys- 
tems of artificial draft are illustrated and described in the author's 
Steam Boilers. 



300. Table Showing Height Of Water Column Due To 
Unbalanced Pressures In Chimney 100 Feet High. Tempera- 
tures are in degrees Fahrenheit. 



Temperature 


Temperature of external air. (Barometer 14.7 lb.) 


in 
chimney 


0° 


10° 


20° 


30° 40° 


50° 


60° 


70° 


80° 


90° 


100° 


200° 


453 


.419 


.384 


.353 


.321 


.292 


.263 


.234 


.209 


.182 


.157 


220 


.488 


.453 


.419 


.388 


.355 


.326 


.298 


.269 


.244 


.217 


.192 


240 


.520 


.488 


.451 


.421 


.388 


.359 


.330 


.301 


.276 


.250 


.225 


260 


.555 


.528 


.484 


.453 


.420 


.392 


.363 


.334 


.309 


.282 


.257 


280 


.584 


.549 


.515 


.482 


.451 


.422 


.394 


.365 


.340 


.313 


.288 


300 


.611 


.576 


.541 


.511 


.478 


.449 


.420 


.392 


.367 


.340 


.315 


320 


.637 


.603 


.568 


.538 


.505 


.476 


.447 


.419 


.394 


.367 


.342 


340 


.662 


.638 


.593 


.563 


.530 


.501 


.472 


.443 


.419 


.392 


.367 


360 


.687 


.653 


.618 


.588 


.555 


.526 


.497 


.468 


.444 


.417 


.392 


380 


.710 


.676 


.641 


.61l|.578 


.549 


.520 


.492 


.467 


.440 


.415 


400 


.732 


.697 


.662 


.632 


.598 


.570 


.541 


.513 


.488 


.461 


.436 


420 


.753 


.718 


.684 


.653 


.620 


.591 


.563 


.534 


.509 


.482 


.457 


440 


.774 


.739 


.705 


.674 


.641 


.612 


.584 


.555 


.530 


.503 


.478 


460 


.793 


.758 


.724 


.694 


.660 


.632 


.603 .574 


.549 


.522J.497 


480 


.810 


.776 


.741 


.710 


.678 


.649 


.620 .591 


.566 


.540 


.515 


500 


.829 


.791 


.760 


.730 


.697 


.669 


.639 


.610 


.586 


.559 


.534 



264 



STEAM POWER PLANT AUXILIARIES 



[Div. 8 



301. Table Of Actual Temperatures Obtained in Typical 
Economizer Installation (Green Economizer Co.), 



Name of plant 



Temperatures 



Of flue gases 



Entering 
econo- 
mizer, 
deg. 
fahr. 



Leaving 
econo- 
mizer, 
deg. 
fahr. 



Of water 



Entering 
econo- 
mizer, 
deg. 
fahr. 



Leaving 
econo- 
mizer, 
deg. 
fahr. 



Hollister Mining Company 

Mac Sim Bar Paper Co 

Mary Charlotte Mining Co 

Kellogg Toasted Corn Flake Co 

Louisville Water Works 

Wessniger-Gaulbert Realty Co 

Galveston Ice Company 

Gilbert Paper Company 

Bemis Bros. Bag Co 

Great Northern Railway 

Portland Railway & Light Co 

Graniteville Manufacturing Co 

Arkwright Mills 

Arnold Print Works 

Blackstone Manufacturing Co 

Champion International Co 

Granite Mills No. 1 

Hoosic Cotton Mills 

Kunhardt Company 

Lancaster Mills 

Nonotuc Silk Co 

Pierce Manufacturing Co 

Stanley Works 

American Thread Co 

Lonsdale Co 

American Brass Company 

Baltic Mills 

Bridgeport Malleable Iron Co 

Lawton Mills 

Union Metallic Cartridge Co 

Waterbury Clock Company 

American Agricultural Chemical Co 

Chelsea Mills No. 1 

Remington Salt Co 

Saratoga Victory Manufacturing Co 

Delaney & Company 

Bird & Son 

Berlin Mills Co 

Winchester Repeating Arms Co 

Hammermill Paper Co 

Imperial Steel Company 



598 
630 
500 
560 
455 
510 
500 
500 
550 
600 
475 
442 
680 
510 
436 
800 
455 
430 
550 
655 
430 
638 
700 
455 
475 
575 
475 
560 
505 
500 
550 
415 
460 
700 
315 
670 
600 
540 
553 
600 
386 



298 
418 
300 
320 
325 
325 
250 
320 
350 
420 
245 
263 
375 
280 
263 
570 
245 
275 
300 
270 
325 
434 
300 
345 
234 
425 
265 
300 
301 
350 
330 
265 
299 
500 
225 
450 
300 
217 
312 
300 
265 



202 
190 
210 
212 
164 
192 
100 
212 
180 
140 
130 
118 
120 
118 

77 
240 

82 
122 
100 
110 

36 
121 
150 
150 
101 
160 
120 
206 
140 
175 
130 

64 
160 
160 

75 
180 
130 

78 
194 
150 

95 



306 
292 
310 
306 
279 
334 
222 
310 
280 
250 
220 
230 
250 
238 
196 
332 
180 
232 
230 
230 
174 
240 
330 
270 
208 
260 
274 
325 
262 
300 
260 
245 
270 
300 
190 
310 
260 
230 
290 
270 
230 



Sec. 302] 



FUEL ECONOMIZERS 



265 



302. The Relative Current-Flow Of The Water And Gases 
Passing Through An Economizer may be: (1) In the same 
direction. (2) In opposite directions. The first is called a 
parallel-flow. The second is called a contra-flow. For a 
parallel-flow (Fig. 270) the feed-water and the combustion- 
gases enter the economizer at the same end. Thus the coolest 
part of the water-current abstracts heat from the hottest part 
of the gas-current. For a contra-flow (Fig. 271) the feed- 
water and the gases enter the economizer at opposite ends. 
Thus the water is first heated by the cooler gases and later, as 
it passes on through the economizer, by the hotter gases. A 
larger transfer of heat from the gases to the water occurs with 



Feed-Water Outlet-^ 



r. n c ^ /^ rs o, o> 
from* 
-Boiler. 




y - Feed -Water Inlet 



Fig. 270. — Illustration Of Water 
And Gas Flow In Parallel Flow Econ- 
omizer Installation. 



-Feed-Wafer Outlet 

r^\ r^ r^\ <c\ ^ ^W tos 
to 




Fig. 271.— The Flow Of Water And 
Gases In Counter Flow Economizer. 



a contra-flow than with a parallel-flow. This is due to the 
minimum temperature-difference, between the water and the 
gases, being (Fig. 272) greater with a contra-flow than (Fig. 
273) with a parallel-flow. 



Explanation. — Suppose the temperatures of the combustion-gases 
and feed-water, at entrance to an economizer which is arranged for a 
parallel-flow, are, respectively, 600 deg. fahr. and 100 deg. fahr. Also, 
suppose the temperature of the gases at exit from the economizer to be 
340 deg. fahr. Then (Fig. 273), if the water is to be heated to a tempera- 
ture of 220 deg. fahr., the economizer must have about 8,000 sq. ft. of 
heating-surface. The minimum temperature difference = 340 — 220 = 
120 deg. fahr. 

Now suppose the economizer to be arranged for a contra-flow. Then 
(Fig. 272), 7,300 sq. ft. of heating-surface would suffice to produce a final 



266 



STEAM POWER PLANT AUXILIARIES 



[Div. 8 



feed-water temperature of 220 deg. fahr., with a gas-temperature, at 
exit, of 340 deg. fahr. The minimum temperature difference 
= 340 - 100 = 204 deg. fahr. 



m 



200 



^ 1 1 1 1 1 1 1 1 1 1 1 1 1 


\ 1 1 1 1 1 1 II 1 1 1 1 1 








■ \uJ-i- 


ct 




s^ 


Sy 


s> 


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^s*i 


J\ 


s 


J 




T 


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v^ i 


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*"S»» Js 


x"*"** 






"t 


T 



S 1 II 1 1 1 II II 1 II 


\ 1 1 1 1 II 1 1 1 1 1 | 1 


--^- Gases Enter at 600° Fahr. ~ 


,„ \ -Water Enters at 100 Fahr. 


\ 


*v i 


" *&T_ 


^&- 


X 400 ^ 


4U0 s 


s ^ it 


^i 




v. 300 > 




* -t 




O i _ _. 




' w " '^^""^ T 


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<U ^ 




W | 


V. 1 


w» X 


X 


* fl - ± 



" v mo mo 3000 mo sooo mo wo woo sooo mm 

Scj. Ft. of Economizer Surface 

Fig. 272. — Characteristics Of Econ- 
omizer Arranged For Contra-Flow Of 
Gases And Water. (Power, Apr. 22, 
1919.) 



" 1000 2000 3000 4000 5000 6000 7000 8000 3000 lOOQl 

Sq.Ft.of Economi zerSurf wee 

Fig. 273. — Characteristics of Econ- 
omizer When Gases And Water Flow 
In The Same Direction. (Power, Apr. 
22, 1919.) 



303. The Ratio Of The Loss Of Combustion-Gas Tempera- 
ture To Gain Of Feed-Water Temperature In An Economizer 

may be found by the following formula: 

C W 

(82) X=T f§ + T fw = -p^ (ratio) 

Wherein X = ratio of decrease of gas-temperature to increase 
of water-temperature. T/ g = loss of gas-temperature, in 
degrees Fahrenheit. Tf w = gain of water-temperature, in 
degrees Fahrenheit. C w = assumed specific heat of water = 
1.0. C g = assumed specific heat of combustion-gases = 0.24. 
W w = weight of water evaporated in boiler, in pounds per 
pound of coal burned. W g = weight of combustion-gases, in 
pounds per pound of coal burned. 

Example. — An average of 13 lb. of combustion-gases are produced, 
per pound of coal burned, in a boiler-furnace. An average of 6.5 lb. 
of water, per pound of coal burned, is evaporated in the boiler. What 
is the ratio between the gas- and water-temperature changes which occur 
in the economizer? Solution.— By For. (82) X = T fa /T fw =C W W W / 
C g W g = (1 X 6.5) -T- (0.24 X 13) = 2.08. 



Sec. 304] 



FUEL ECONOMIZERS 



2G7 



Note. — The ratio-value obtained in the solution of the preceding 
example is commonly assumed to represent, approximately, economizer- 
performance in general. It is based upon the assumption that infiltra- 
tion of air through the setting (Sec. 297), and radiation of heat from 
the economizer, are both reduced to a minimum. 



304. Additional Heat- 
ing-Surface Obtained By 
Installing An Economizer 
Will Prove More Effect- 
ive, Than Would An 
Equivalent Addition Of 
Boiler Heating-Surface, 
in absorbing heat from 
the combustion -gases 
which are discharged from 
a set of boilers. This is 
due to the greater temper- 
ature-difference between 
the water and the gases. 




Fig. 274. — Diagram Showing Gas Tempera- 
tures In A Water-Tube Boiler Containing 6,000 
Sq. Ft. Of Heating-Surface. 



Explanation. — Suppose the steam-pressure in a boiler is 150 lb. per 
sq. in., gage. Then the temperature of the water in the boiler will be 
about 358 deg. fahr. Suppose the boiler heating-surface is of such 
extent that it will lower the combustion-gas temperature in the last 

pass (Figs. 274 and 275) to 500 
deg. fahr. Then the temperature- 
difference between the inside and 
the outside of the boiler heating- 
surface = 500 - 358 = 142 deg. 
fahr. If the gases were now to flow 
in contact with additional boiler 
heating-surface, the rapidity of heat- 
transfer from the gases to the water 
would (see the author's Practical 
Heat) be in direct proportion to 
this temperature-difference. But if 



if 

<ns2Q< 

.-> .j. 600 



c fe 400 



1 




































































































































































































































































^ — 


1 |cj 


tp» 


W - 




~< 


:-.= 




l^Vy^L? 


: 


z 


3 


: - 


1 s 


: t 


: l 


i s 


9 


: :: 



Per Cent of Boiler Tube Surface Passed 
Over 



Fig. 275. — Graph Showing Gas Tem- 
peratures In A Water-Tube Boiler Oper- 
ating At About 10 Sq. Ft. Of Heating the gases were to traverse an econ- 
Surface Per Boiler Horse Power. (Gas om i zer through which water at 150 

SrSecL'r ^ 1S ° 0mPar " deg. fahr. i. being pumped, then the 

rapidity of heat-transfer from the 
gases to the water would be in direct proportion to a temperature- 
difference of 500 - 150 = 350 deg. fahr. 



268 



STEAM POWER PLANT AUXILIARIES 



[Div. 8 



305. The Least Temperature -Difference That Can Be 
Profitably Permitted Between The Inside And The Outside 
Of Boiler Heating-Surface may be determined, approximately, 
from the chart (Fig. 276) which has been computed for average 

conditions of boiler service. From 
the temperature-difference so ob- 
tained, the lowest flue-gas temper- 
ature may be computed. The 
maximum amount of heating- 
surface which a boiler should have, 
for given conditions of operation, 
may then (Fig. 277) be determined. 




D 



1.00 , 2.00 3.00 100 
Cost of Coal in Dollars per Ton 

Fig. 276.— Chart Showing The 
Least Temperature Difference Be- 
tween The Temperatures Of Gases 
And Water Under Different Con- 
ditions At Which Additional Boiler 
Surface Ceases To Pay Dividends. 
(Green Economizer Co.) 



Note. — If the heating-surface of a 
boiler is too extensive, the temperature- 
difference in the last pass will be insuf- 
ficient to insure effective heat-transfer. 
Example. — A set of boilers is to deliver 
steam at a pressure of 200 lb. per sq. in. The daily period of operation 
is to be 12 hr. Coal will cost $3 per ton. What is the maximum 
amount of heating surface, consistent with economical performance, 
which each boiler should have? 



£ 100 



s 500 



£300, 





























































































• 


































• 




• • 


•• 


••. 


































• 




> 
• 


• 


•,. 
































• 


/ 


» 




• 


• 


( 


• 


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• 


s 


• 




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• 


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• 


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•. 












































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12 



*~ Boiler Surface in Square Feet per Boiler Horse Power Developed 

Fig. 277. — Chart Showing Flue Gas Temperatures Corresponding To Different 
Amounts Of Heating Surface Per Boiler Horse Power Developed. Each Point Repre- 
sents A Test. (Green Economizer Co.) 



Solution. — For a 12-hr. daily run, with coal at $3 per ton, the least 
temperature-difference, consistent with profitable operation of the boiler, 
is (Fig. 276) 200 deg. fahr. The temperature of steam at a pressure of 
200 lb. per sq. in., gage, is about 388 deg. fahr. 



Sec. 306] 



FUEL ECONOMIZERS 



269 



perature of the combustion-gases would be (200 + 388) = 588 deg. fahr. 
The permissible extent of heating-surface is, therefore, (Fig. 277) about 5.5 
sq. ft. per boiler h.p. 

306. The Least Temperature -Difference That Can Be 
Profitably Permitted Between The Inside And The Outside 
Of Economizer Heating-Surface may be determined, ap- 
proximately, from the chart (Fig. 278) which has been com- 
puted for different conditions of 
economizer service. Local con- 
ditions, peculiar to individual 
plants may, however, sometimes 
affect the accuracy of the deter- 
minations. 




1.00 2.00 3.00 4.00 
■^ Cost of Coal in Dollar s Per Ton 

Fig. 278.— Chart Showing The 
Least Temperature Differences Be- 
tween The Temperature Of Gases 
And Water, Under Different Con- 
ditions, At Which Additional Econ- 
omizer Surface Ceases To Pay- 
Dividends. (Green Economizer 
Co.) 



Example. — A contra-flow economizer 
(Fig. 271) is to be installed in connection 
with the set of boilers mentioned (Sec. 
305) in the preceding example. The tem- 
perature of the feed-water, at entrance to 
the economizer, is to be 150 deg. fahr. 
The temperature of the entering gases is 
588 deg. fahr. What should be the least 
temperature, consistent with economy, 
of the gases issuing from the economizer? 

Solution. — For a 12-hr. daily run, with coal at S3 per ton, the least 
temperature-difference, consistent with profitable operation of the econo- 
mizer, is (Fig. 278) 90 deg. fahr. Hence, the exit-temperature of the 
gases shoidd le (90 + 150) = 240 deg. fahr. 

307. The Ratio Of Economizer Heating-Surface In Square 
Feet To Boiler-Horsepower usually ranges from about 4:1 
to 8:1. An extent of heating-surface, for the economizer in 
the above example, which would give a mean between these 
ratios, would probably cool the gases from the entering tem- 
perature of 588 deg. fahr. to the requisite 240 deg. fahr. at 
exit. 

Example. — The Commonwealth Edison Co., Fisk St. Station has a 
nominal boiler horse power of 1,225 per unit and an economizer surface 
of 8,500 sq. ft. per unit, or a ratio of (8,500 ■£■ 1,225) = approximately 7.1. 

308. The Rate Of Heat -Transfer Between The Combustion- 
Gases And The Water In An Economizer is mainly conditional 
upon the rate of the gas-flow through the economizer. It may 



270 STEAM POWER PLANT AUXILIARIES [Div. 8 

range from about 1.5 to 5.5 B.t.u. per hr. per sq. ft. of heating 
surface per deg. of temperature-difference between the gases 
and the water. An average figure for the rate of heat transfer 
in good modern economizer installations is about 4.3 B.t.u. 
per hr. per sq. ft. per degree difference in temperature. It is 
assumed in this statement, that the heating-surface is clean, 
and that there is no air infiltration through the setting. It is 
also presumed that the flow-velocity of the water is uniform 
for various rates of heat-transfer. 

Note. — The minimum rate of heat-transfer, noted above, may occur 
with a gas-flow of about 1,300 lb. per hr. per sq. ft. of heating-surface. 
The maximum rate may occur with a unit gas-flow of about 5,000 lb. 
per hr. per sq. ft. of heating-surface. Intermediate rates of heat-transfer 
and gas-flow would be approximately proportional to these values. 

Example . — An economizer is required to raise the temperature of 75,000 
lb. of water per hr. through 15 deg. fahr. The average temperature- 
difference between the combustion-gases and the water is assumed to be 
200 deg. fahr. The rate of heat-transfer is assumed to be 4 B.t.u. per 
hr. per sq. ft. of heating-surface per deg. of temperature-difference be- 
tween the gases and the water. The assumed specific heat of the water = 
1.0. What is the requisite area of heating-surface? 

Solution. — The hourly transfer of heat per sq. ft. of heating surface = 
(4 X 200) = 800 B.t.u. For a temperature-rise of 15 deg. fahr., each 
pound of water must absorb (15 X 1.0) = 15 B.t.u. Therefore, the 
requisite heating-area = (75,000 X 15) -f- 800 = 1,406 sq. ft. 

309. The Percentage Of Fuel-Saving Due To Economizer- 
Service (Fig. 279) may be computed by the following formula : 

, M x v 10 °(g% -T' fw ) , ,. 

(83) X = -— -±- (per cent.) 

K J (H + 32) - T' fw 

Wherein X = per cent, of saving. T'/ w — the temperature 
of the feed-water at exit from the economizer, in degrees 
Fahrenheit. T' fw = the temperature of the feed-water at 
entrance to the economizer, in degrees Fahrenheit. H = 
the total heat in the steam, above 32 deg. fahr., in B.t.u. per 
pound. 

Example. — The temperature of the feed-water entering an economizer 
is 160 deg. fahr. The exit-temperature of the water is 305 deg. fahr. 
The boiler-pressure is 200 lb. per sq. in., gage. What is the percentage 
of fuel-saving, due to the economizer service? 



Sec. 310] 



FUEL ECONOMIZERS 



271 



Solution. — The total heat in steam at 200 lb. pressure per sq. in., 
gage, as given in the steam tables = 1199.2 B.t.u. per lb. By For. (83): 
X = 100(7% - 7%)/[(ff + 32) - T' /w ] - 100 X (305 - 160) -J- 
[(1199.2 + 32) - 160] = 13.5 per cent. 

Note. — It is commonly assumed that an economizer will effect a 
fuel-saving of approximately 1 per cent, for each 11-deg. fahr. rise in the 
feed-water temperature. The saving may amount to 25 per cent. 

Example. — A 2,000-horse power boiler plant runs 10 hours per day and 
300 days per year. The coal-consumption is 4.6 lb. per h.p. per hr. The 
coal costs $4 per ton. It is assumed that, with economizer-service, 13 
per cent, of the fuel would be saved. An economizer-installation would 



WW" 


A *■ 


250- 


__:__ __/_u-i 




/J\n „_ 


2.40- 






M.U 1/ \ 

A-A, Economizer J 




" Shut-Off ; T' 
I 








A V -^ /v^ v S 




Average ; f\ i V 


2.10 - 




I 


\ / ! Working 


?00- 


V / 



I 2 3 4 5 6 1 8 9 10 II 12 13 14 15 16 17 
Days of M 



3 19 2Q 21 22 23 24 25 26 27 28 29 30 31 
on t h 



Fig. 279. 



-Diagram Showing Coal Consumption When Economizer Was Stopped And 
When In Operation With Clean Tubes. 

cost $9,500. The annual cost of economizer-operation, -depreciation 
and -repairs would be 15 per cent, of the cost of installation. What 
would be the monetary value of the net yearly fuel-saving, due to the 
economizer-service? What percentage of the cost of the economizer- 
installation would the annual saving represent? 

Solution.— The annual expense for fuel = [(2,000 X 4.6 X 10 X 300) 
-s- 2,000] X 4 = $55,200. The prospective annual cost of economizer 
operation, depreciation and repairs = 9,500 X 15 v 100 = $1,425. 
Hence, the prospective net annual saving with economizer-service = (55,200 
X 13 4- 100) - 1,425 = $5,751. This represents (5,751 -^ 9,500) X 
100 = 60.5 per cent, of the cost of the economizer-installation. 

310. The Increase Of Steam-Generating Efficiency, Due 
To Economizer-Service may vary, according to local condi- 
tions, as follows. 



272 STEAM POWER PLANT AUXILIARIES [Div. 8 

Example. — When the boilers are run slightly above their rated 
capacity, and the heating-surface of the economizer is approximately 
equal to 60 per cent, of that of the boilers, the efficiency-increase may 
be about 6 per cent. When the boilers are run at about their rated 
capacity, with the combustion-gases entering the economizer at a 
temperature of 500 deg. fahr., and with the usual flow-velocity, the 
efficiency increase may be about 5 per cent. 

Example. — When the inherent economy of the boilers is good, due to 
excellence of design, both of boilers and settings, and the boilers are run 
at 200 per cent, of their rated capacity, the efficiency-increase may be 
10 per cent. This percentage of increase is based on a temperature of 
600 deg. fahr. and a flow-velocity of from 1,800 to 2,000 ft. per min. for 
the gases entering the economizer. Also, the economizer heating-surface 
is presumed to equal 70 per cent, of the boiler heating-surface. 

Example. — When the inherent economy of the boilers is poor, due to 
defective design, both of boilers and settings, and the boilers are run at 
less than their rated capacity, the efficiency-increase may be nil; there 
may, under such conditions, be an actual decrease in overall efficiency. 

311. The Principal Advantages Of Economizer Service are : 

(1) Fuel-saving. The saving may amount to from 5 to over 
18 per cent. (2) Increased boiler-efficiency. Where the boilers 
are operated at over 200 per cent, of their nominal rating, and 
the supply of exhaust steam for heating the feed-water is scant, 
due to the auxiliaries being electrically driven, the increase of 
efficiency may be considerable. (3) Diminished contraction 
stresses in steam boilers. This results from the high feed-water 
temperature which is attainable with the economizer. (4) 
Increased flexibility of boiler operation. This results from the 
storage-space which the economizer affords. The large quan- 
tity of hot water in the economizer is instantly available for 
use in the event of a sudden overload. 

312. The Principal Disadvantages Of Economizer Service 
are: (1) Expense for installation. (2) Expense for maintenance. 
This comprises, mainly, the costs of repairing and operating 
(Sec. 294) the soot-scrapers or blowers. (3) Diminished 
overall efficiency of the plant. This may occur where the draft 
is insufficient for economizer operation, or where the boilers 
are operated below their nominal ratings. (4) Bulkiness of the 
requisite equipment. An economizer, and its appurtenances, 
as a motor- or engine-driven draft-fan, requires large floor 
space. If the economizer is erected overhead, much altera- 



Sec. 313] FUEL ECONOMIZERS 273 

tion of piping and structural details will generally be neces- 
sary to make room for the equipment. 

313. The Conditions Which Usually Determine Whether Or 
Not Economizer Should Be Installed are chiefly as follows: 

(1) The total horse power of the plant (Sec. 309). 

(2) The flue-gas temperature. Where the temperature is above pos- 
sibly 450 to 550 deg. fahr., profit may result from using an economizer. 
The higher the temperature, the greater the saving. The flue gases 
should not be cooled below 250 deg. fahr. because the vapor in the gases 
will be condensed on the economizer tubes, especially near the exit. 
This will cause soot to adhere to the surfaces of the tubes. If the coal is 
high in sulphur content, the condensed moisture may collect sulphur 
dioxide from the gases and dilute sulphuric acid result. This corrodes 
the tubes (Sec. 292). 

(3) The boiler pressure. When the pressure is 250 lb. per sq. in., or 
over, an economizer is practically indispensable. 

(4) The character of the load. If the plant is heavily overloaded, either 
steadily or intermittently, there may be need for an economizer. Gener- 
ally a substantial saving may be realized when boilers are operated well 
above their ratings for a large proportion of the time. 

(5) The feed-water temperature. Increased economy may result from 
the high temperature which may be obtained with an economizer. A 
great saving should result in plants running condensing when motor- 
driven auxiliaries are employed. Under these conditions there is usu- 
ally insufficient exhaust steam to heat the feed-water. The economizer 
should deliver the water at a much higher temperature — much higher 
than 210 deg. fahr. which is usually the limit for exhaust steam heaters. 
The wider the range over which the economizer heats the water, usually 
the greater is the saving. 

(6) The quantity of exhaust-steam available for heating. When feed- 
water heaters are installed and plenty of exhaust steam is available, 
which would be lost if not used in the heater, an economizer may not 
show any considerable saving. 

(7) The quality of the feed water. If the water contains impurities 
which will form scale in an economizer, then an economizer may prove 
undesirable. 

(8) The available means for supplying sufficient draft, and the cost 
thereof. Lack of building space might necessitate erection of a tall 
chimney. Otherwise, an artificial draft equipment may be necessary. 
The cost of a tall chimney might be prohibitive. Likewise, the expendi- 
ture of from 1 to 4 per cent, of the total power output for driving the 
draft equipment (see author's Steam Boilers) might be prohibitive. 

(9) The initial cost of the economizer. When economizers are required 
to sustain pressures greater than 250 lb. per sq. in., their cost increases 
rapidly with the pressure. 

18 



274 STEAM POWER PLANT AUXILIARIES [Div. 8 

(10) The price of coal. When the price of coal is high there is more 
saving than when it is cheap, unless the cost of the economizer and its 
operation also are high in the same proportion. 

314. Economizers Should Be Inspected Periodically. 

There should be a monthly overall detail inspection. Certain 
elements may require more frequent attention. Inspection 
should cover the following details : 

(1) The external surfaces. Leaks and soot-deposits should be looked 
for. Soot should be removed. Leaks should be stopped. They cause 
corrosion and tend to produce soot- and rust-scale on the tubes. 

(2) The internal surfaces. A scale-forming tendency (Sec. 295) in 
the tubes should be looked for. If such exists, the economizer should be 
opened as frequently as practicable and the tubes washed with a hose. 

(3) The safety-valve. Corrosion between the valve and seat should be 
looked for. Also, the valve mechanism should be tested for freedom of 
movement. 

(4) The blow-off valves. The valve discs should be examined. Like- 
wise the packing of the stems. If the packing is dry and hard, the stems 
should be repacked. The stems should work freely. 

(5) The flange- joints. Leaking joints should be repacked. Also, any 
straining effect on the joints, due to restraint of expansion and contrac- 
tion in the pipe-lines, should be rectified. 

(6) The soot-scrapers or blowers. The distance traveled by the scrapers 
should be noted. It should be the full length of the tubes. The blower- 
nozzles should be examined. If the nozzle-orifices have been enlarged 
by erosion, the nozzles should be renewed. 

(7) The gearing and reversing-mechanism of the soot-scrapers. Broken 
gear-teeth should be looked for. The security of bolts, pins and cotters 
should be tested. The lubrication of the bearings should be noted. 

(8) The dampers. The devices for damper-adjustment should be 
tested. 

(9) The setting. Search should be made for cracks in the masonry. 
Leakage of air around door-frames should be looked for. Suspected 
places may be tested by applying the flame of a candle or torch, with the 
stack damper open. 

(10) The soot-pits or chambers. These should be entirely emptied of 
their contents as frequently as is necessary. 

(11) The thermometers. The accuracy of the instruments should be 
noted. 

Note. — Examinations For Evidences Of Pitting On The Interior 
Surfaces Of Economizers (see the author's Steam Boilers) should be 
made annually. Facility in making these inspections may sometimes 
necessitate a partial disassembling of the economizer sections. 



Sec. 315] FUEL ECONOMIZERS 275 

315. The Cost Of An Economizer And Its Installation is 

usually computed on the ratio of the economizer heating- 
surface to the boiler horse power. When this ratio is 5:1 the 
cost, prior to the Great War, was about $6 per boiler horse 
power. When the ratio is 4.8:1, the prewar cost, for plants 
containing 1000 boiler horse power, or more, was about S5.50 
per boiler horse power. Otherwise, the cost, regardless of 
either the size of the installation or the ratio mentioned above, 
may be based directly on the extent of economizer heating- 
surface. On this basis, a prewar cost of $1.20 per sq. ft. was 
usually assumed. Power Plant Engineering, Apr. 1, 1920, 
states that cost, including fan, motors, etc., will now average 
about $4 per sq. ft. of economizer surface. 

QUESTIONS ON DIVISION 8 

1. What is the function of a fuel-economizer? 

2. Describe an independent fuel-economizer. An integral fuel-economizer. 

3. What materials are used in economizer construction? What are the relative 
advantages and disadvantages of these materials? 

4. What detrimental effects may result from coatings of soot on economizer surfaces? 

5. In what manner do sedimental deposits in economizer-tubes affect the economy 
of the apparatus? 

6. What physical injury may result from impurities in the water pumped through 
an economizer? 

7. Describe the operation of economizer tube scrapers. 

8. How may mineral substances in the feed-water be prevented from forming scale 
in economizer-tubes? 

9. Why does not hard scale form as readily in an economizer as in a boiler? 

10. Explain the ill-effects of air-infiltration through an economizer-setting. 

11. How may air-infiltration through an economizer-setting be detected? 

12. How does air-infiltration affect the quality of the combustion-gases in an econo- 
mizer? What may be regarded as a reasonable drop in the percentage of CO2 in the 



13. If a boiler plant is being operated with natural draft, what would be the probable 
effect on the draft if an economizer were installed? 

14. What is the usual method of supplying draft for boiler plants which are equipped 
with economizers? 

15. Enumerate the principal advantages of economizer service. The principal 
disadvantages. 

16. Enumerate the conditions of boiler-service which chiefly determine the advis- 
ability of installing an economizer. 

17. Enumerate the structural details and service-conditions toward which economiz- 
er-inspections should be particularly directed. 

18. Explain why economizer heating-surface is more effective in absorbing heat from 
the combustion-gases leaving a boiler than an extension of boiler heating-surface would 
be. 

19. What is a contra flow in an economizer installation? A parallel flow? Which is 
the more effective? 

20. What is the average ratio, in economizer operation in general, of the drop of 
combustion-gas temperature to the rise of feed-water temperature? Give some ex- 
amples, approximating this ratio, as observed in typical economizer-installations- 



276 STEAM POWER PLANT AUXILIARIES [Div, 8 

21. What service-condition mainly determines the extent of economizer heating- 
surface that can be profitably used? 

22. Upon what service-condition is the rate of heat-transfer in an economizer mainly 
contingent? 

23. What percentages of increase of steam-making efficiency, for various conditions of 
boiler-service, may ordinarily be anticipated from economizer-service? 

PROBLEMS ON DIVISION 8 

1. The combustion-gases leaving a boiler have a temperature of 550 deg. fahr. and a 
CO2 content of 12 per cent. On account of leakage of air through the setting, the gases 
leaving the economizer have a temperature of 250 deg. fahr. and a CO2 content of 8 per 
cent. The specific heat of the gases = 0.24. What percentage of the heat is lost when 
the temperature of the outside air is 50 deg. fahr.? 

2. For each 15 lb. of combustion-gases flowing through an economizer there is a 
corresponding water-flow of 8 lb. What should be the ratio of the decrease of combus- 
tion-gas temperature to the increase of feed- water temperature? 

3. A boiler is to deliver steam at a pressure of 175 lb. per sq. in., gage. The tempera- 
ture of steam at this pressure = 377.5 deg. fahr. The boiler will be run 24 hr. per day. 
Coal will cost $3.00 per ton. How much heating-surface, per boiler horsepower, 
should the boiler have in order that it may be operated economically? 

4. It is assumed that an economizer is to be installed in connection with the boiler 
mentioned in Problem 3. It is also assumed that the feed-water will enter the economiz- 
er at a temperature of 200 deg. fahr. What should be the least temperature, consistent 
with economical operation, of the combustion gases at exit from the economizer? 

5. The average temperature-difference between the feed-water and the combustion- 
gases in an economizer is assumed to be 300 deg. fahr. The economizer is required to 
raise the temperature of 50,000 lb. of water per hour through 50 deg. fahr. The rate of 
heat transfer is assumed to be 5.5 B.t.u. per hr. per sq. ft. of heating surface per degree 
of temperature-difference between the gases and the water. The assumed specific heat 
of the water = 1.0. What should be the area of the economizer heating-surface? 

6. The temperature of the water entering an economizer is 110 deg. fahr. The 
temperature of the water at exit is 250 deg. fahr. The boiler-pressure is 175 lb. per sq. 
in., gage. The total heat, above 32 deg. fahr., in steam at this pressure = 1197.3 B.t.u. 
per lb. What is the percentage of fuel-saving? 

7. A 2400-horse power boiler plant runs 24 hr. per day and 300 days per year. The 
coal-consumption is 4.3 lb. per h.p. per hr. The coal costs $4.25 per ton. It would 
cost $12,000 to install an economizer in this plant. The annual cost of operation, 
maintenance and depreciation would amount to 15 per cent, of the cost of installation. 
Assuming that the economizer would effect a fuel-saving of 12.3 per cent., what would 
be the monetary value of the saving per year? 



DIVISION 9 



STEAM CONDENSERS 



316. A Steam Condenser As Used In Connection With A 
Steam Engine is a device for reducing exhaust steam to water. 
The purpose of a condenser is to increase the power which is 
developed by an engine from a given quantity of steam; or, 
conversely, to decrease its steam consumption for a given 
power output. 

Note. — A Condenser Increases Engine Economy By Creating A 
Partial Vacuum in a container into which the engine discharges its 
exhaust steam. The method of creating the vacuum is to cool the ex- 
haust steam sufficiently so that 
it will condense to water, which 
occupies very much less space. 
The effect of the partial vacuum 
thus created is to give the engine 
10 to 15 lb. per sq. in. more work- 
ing pressure without any in- 
crease in boiler pressure or 
material increase in fuel con- 
sumed. Greater working pres- 
sure with a given quantity of 
steam results in greater power 
output. 




317. How A Condenser 
Increases The Working 
Pressure of a steam cyl- 
inder may be demonstrated 
thus: — Fig. 280 shows two 
elementary steam cylinders 
surrounded by air at nor- 
mal atmospheric pressure. This pressure is equal to about 14.7 
lb. per sq. in. at sea level. That is, any object exposed to the 
air at sea level has 14.7-lb.-per-sq.-in. pressure exerted on it 
from all directions. Assume that a pressure of 100 lb. per sq. 

277 



Fig. 280.— Showing The Effect Of Vacuum 
Produced By Condensation On The Working 
Pressure Of A Steam Piston Having An Area 
Of 1 Sq. In. 



278 



STEAM POWER PLANT AUXILIARIES 



[Div. 9 



in. as indicated by a steam gage, is exerted on the under side of 
a piston, A, of one square inch area. Then the piston will 
have a lifting force of 100 lb. There is really 114.7 lb. pressure, 
pushing on the under side and 14.7 lb. on the upper, but only 
the difference, 100 lb. is effective. But, suppose the space 
above piston, B, which has the same area, is first filled with 
steam and then condensed (assuming that it will condense 
completely). Then there will be no pressure on the upper 
side of the piston and the whole 114.7 lb. on the lower side 
becomes effective. The piston then has a lifting force of 
114.7 lb. Thus, by condensing the steam above the piston in 
B its lifting force is increased by 14.7 lb. 

318. Power Was Developed By Condensation in Primitive 
Steam Engines. — The primitive engine of Newcomen shown in 
conctensing Fig. 281 works entirely by con- 



■Piston 




... Water 

\i\ supply^ densation. 



Explanation. — On the upstroke 
steam, at a little above atmospheric 
pressure, flows into the cylinder, C, 
through the valve, /, and is then 
condensed by a jet of cold water, S, 
admitted at V. The partial vacuum 
thus created allows the atmospheric 
pressure above to force the piston, 
P, down so that power is developed 
on the downstroke. The condensed 
steam and condensing-water are 
drained out through valve, 0, while 
the piston is on the upstroke. A 
weight, W, counterbalances the piston 
and is connected to a pump rod, H, 
which does the work of the engine. By alternately admitting steam 
into C and condensing it therein, rod, H, is forced to move up and down 
and thereby pump water. 

319. An Improvement Of Newcomen's Engine Was Made 
By Watt (Fig. 282) who condensed the steam in a separate 
chamber or condenser, D, with a jet of cold water. The chief 
advantage of this arrangement over the former is that in this 
the cylinder remains hot and the condenser cold so that neither 
has to be alternately heated and cooled, as with Newcomen's 
arrangement. Watt then made his engine double-acting 



Fig. 281. — Newcomen's Condensation 
Engine. (Year 1763.) 



Sec. 3201 



STEAM CONDENSERS 



279 



(Fig. 283) and operated the valves, /, J, and Q, by a 
system of levers. The condenser, D, then acted continuously 
and the condensed water had to be pumped out of the conden- 
ser. Small amounts of air also accumulated in the condenser 
and were pumped out. 



Condzns 
Ung Wotr 




Fig. 282. — Watt's Condensation Engine. 





Condensing 






Water-. ' 


t 






?-z-z-z- 


--,■ 

3 


Steam 






ssa 


t^-CTv^ 


Inlets 




1 

| 




i m 


% 








T 


fm 


R 




^jj^ji{jl 








— ■ 


_- ~=^x. C: -^J 




> 








Condenser. 




% 






D 


Water 




1 








Y.tj?^A?ii^i i :^A 


VM&^M/M&m 


. 283. — Watt's Double-Acting Con 




< 


iensing Engine. 





320. How The Condenser Saves Steam may be shown by 
either of two methods: (1) By comparing the thermal effi- 
ciency of condensing and non-condensing operation (Sec. 321). 
(2) By computing the ratio of mean effective pressures for con- 
densing and non-condensing operation (Sec. 322). The first 
method is based on the work done by the steam in expanding 
and is therefore, not accurately applicable to those slide-valve 
and similar engines in which the expansion is either zero or 
small. The ideal conditions which this method assumes are 
approached in the compound Corliss engine and the steam 
turbine. The second method gives a fair estimate of the actual 
power increase effected by the condenser except that the power 
required by the condenser auxiliaries must be deducted. 

Note. — The relation of the power developed by a condensing engine 
to that by a corresponding non-condensing engine is shown in Fig. 284. 
The area DCFE represents the increase in power due to the condenser. 
This may be compared to the power developed by the primitive con- 
densation engine (Fig. 281). The area GABCD represents the power 
developed by the engine when operating non-condensing. The modern 



280 



STEAM POWER PLANT AUXILIARIES 



[Div. 9 



condensing engine develops the sum of these amounts of power from the 
same quantity of steam, as represented by the total area GABEF. 







'F Work Done During Non-Condensing Stroke- 






01/3 q, ---.-.. A A p om p ur i n g cnntfawny ?trofa± | 1 1 1 1 1 










J5Q -----\ -+---_ Saving In Back Pressure=FD=12Lb/5a.In. 






f X Condenser. 




t\ \x 


v+ 


3± 


















& i \ 1 f ' 




1"tft "I -i- ^v. 




3 ft ] | ^Sv 1 




a 1 p t|lff^[4UU MJM|MJ|I 1 


"* n:_.k._...j£ i ~t T -- Ir" 


U R "* — H 1 H Q 


::^=S===±^-===-==S===i=======ffli==-====E£:: 



«__ Length Of Stroke -X 

Fig. 284. — Theoretical Indicator Cards Showing Work Gained By Condensing Over 
Non-Condensing Operation. (Gain in work of condensing over non-condensing 
operation =33 per cent.) 



___ — r . 




..it IE ± 


-j-rArecr Ot Non-Condensing Card-GABCDG' 






K Ai 




?on i' i** ^ 




lvu g , -jr 




- it it 


Done Durina condensing Stroke. 1 1 


■C u Jt 




° t i i 




£ + i 


sent Tne Same Work Done. Win? Conaen- 


^ t 




?4- \ 




£ L ' 


-W/'rh Non-Condensing At A. The Additional 


1 L 








«$» ° .- ' 


.indicates The Greater Steam Consumption. 


^ E k. 




x 4 A 




-R t k. 




- 1 : :$ $ 








c -- 5s,_ S ~3E ~ : 




• ^ S,- . 








? inn 5 !* 




h luu IS s, 




3 ± s S r " 




w > ^i 




«« 5 V 




g - - si - 








°- , S r : 




s 






-^ s> * t " i 


| 50 ±__. 




o 


™EEEE~!==3»^~::=!-=::==]EEEE 












^ -v ~ ?• 


i' 1 — ! 11 1 1 1 1 1 M 1 1 M Ml 


MM +- J — 4--^ — i : — — 'c , 








1 1 7i r?^ iJ- 


o D-+ff+H4-hH4-H4-l44-H4 


■4— 1— r-H-Fi ...,--4.fUHC !■ - 



K Length Of Stroke •- -H 

Fig. 285. — Theoretical Indicator Cards Showing Difference In Steam Consumption Of 
Condensing And Non-Condensing Operation For Equal Work Area. 

When steam is used expansively (as indicated by the curved expansion 
line AB) a given difference in pressure below the atmospheric line 
(such as P 3 , Fig. 284) represents much more power than a similar differ- 



Sec. 321] STEAM CONDENSERS 281 

ence in pressure above the boiler pressure, P 4 . In other words, 13 lb. per 
sq.in. of vacuum in a condenser increases the power of a good engine much 
more than 13 lb. per sq. in. more boiler pressure. 

Note. — The areas in the indicator cards (Figs. 284 and 285) represent 
energy or work delivered during one stroke of an engine but, assuming 
a constant engine speed, they also represent the proportional power 
developed. 

321. The Increase In Thermal Efficiency Effected By A 
Condenser may be estimated by the steam temperature 
relations. The greatest possible thermal efficiency of any 
heat engine is represented by the equation: 

(84) E< = Tl ~ T2 (decimal) 

Ti 

The efficiency which this formula gives may be approximated 
by an actual engine but can never be attained. It might be 
attained only by an ideally-perfect engine. (See the author's 
Practical Heat) Wherein: E t = the greatest possible ther- 
mal efficiency of any heat engine. Ti = the absolute tempera- 
ture at which the steam is admitted. T 2 = the absolute 
temperature at which the steam is exhausted. Absolute 
temperature = 460 deg. + the temperature in deg. fahr. 

Example. — Assume an engine using saturated steam at 115 lb. per 
sq. in. abs. As shown by a steam table this steam has a temperature of 
338 deg. fahr. or 338 + 460 = 798 deg. fahr. abs. It exhausts, when 
running non-condensing, at a pressure of 16 lb. per sq. in. abs.; and, 
when running condensing at 2 lb. per sq. in. abs. These pressures, as 
shown by a steam table, correspond to 676 and 586 deg. abs. respectively. 
Hence, by For. (84), the ideal efficiency non-condensing is: E t = (Ti — 
T 2 )/Ti = (798 - 676) /798 = 15.3 per cent. The ideal efficiency con- 
densing = E t = (798 - 586) /798 = 26.6 per cent. 

322. The Theoretical Saving In Power Due To The Use Of 
A Condenser may be computed by the following formula: 

(85) Saving = — ^^ (per cent.) 

Wherein: P hm v = the vacuum obtained in the condenser, in 
inches of mercury. P m = the mean effective pressure of the 
engine running non-condensing, in lb. per sq. in. 

Note. — The saving is much more than proportional to the increase in 
working pressure of the engine. That is (Fig. 284) the saving is propor- 
tional to P 3 -5- Pi not to Pz -r- P\. The mean effective pressure is found 



282 



STEAM POWER PLANT AUXILIARIES 



[Div. 9 



in practice by measuring the area of an actual indicator diagram with a 
planimeter and dividing this area by the length of the diagram. If this 
value (represented by P 2 , Fig. 284) is multiplied by the constant of the 
spring used, (for instance 80 for an 80-lb. spring) the mean effective pres- 
sure in pounds per square inch, P m , For. (85), will be obtained. The 
pressure difference due to the condenser (P 3 , Fig. 284) applies evenly 
throughout the stroke and so the vacuum obtained in the condenser 
may be taken as proportional to P 3 if reduced also to pounds per square 
inch. The vacuum expressed in inches of mercury may be reduced to 
pounds per square inch by dividing by 2.03. The saving, then = 

(per cent.) 



(86) 



100 X 



100 X 



X 2.03 P m 

Example. — The boiler pres- 
sure of an engine is 100 lb. per 
sq. in. gage and the mean effec- 
tive pressure of the engine run- 
ning non-condensing is 44.6 lb. 
per sq. in. What theoretical sav- 
ing will result from condensing 
operation in a 26 in. vacuum? 

Solution. — By For. (85), the 
saving = 49 P hm v/P m = 49 X 
26/44.6, 28.6 per cent. 

Note. — Fig. 285 shows the 
effect of the condenser on en- 
gine economy keeping the 
amount of power developed 
constant, instead of keeping 
the amount of steam used con- 
stant as on Fig. 284. 

323. The Steam Saving 
Due To A Condenser On 
The Basis Of Decreased 
Back Pressure may be de- 
termined very closely by 
the use of a suitable graph 
(Fig. 286) . First the non- 
condensing steam consump- 
tion is determined from the 
graph, which gives values for an ideal engine. Similarly the 
ideal condensing consumption is determined. Then the ratio 
of these values is applied to the actual steam consumption 
considered. 




20 40 60 80 100 120 140 160 180 200 
Initial S+eam Pressure Absolute 
Lb. Per Sq. In 

Fig. 286. — Diagram Showing The Steam 
Consumption Of A Perfect Steam Engine When 
Receiving Steam At Different Pressures And 
Exhausting Against Different Back Pressures. 
(Note bad effect of high back pressure. The 
steam consumption of actual engines is affected 
in about the same proportion.) (Harrison 
Safety Boiler Works Catalog.) 



Sec. 324] 



STEAM CONDENSERS 



283 



Example. — Consider an engine working at 100 lb. per sq. in. abs. with 
1 lb. per sq. in. gage back pressure, consuming 25 lb. of steam per h.p. 
hr. How much steam will it use if operated condensing with a 26 in. 
vacuum? 

Solution. — Locate ordinate A (Fig. 286) corresponding to 100 lb. per 
sq. in. abs. on the lower scale and on this ordinate, find B and C corre- 
sponding to 1 lb. gage per sq. in. and 26 in. of vacuum. The correspond- 
ing ideal consumptions are about 19 and 10 lb. per h.p. hr. That is, the 
ideal condensing consumption is 10/19 of the ideal non-condensing con- 
sumption. But the engine actually consumes 25 lb. per h.p. hr. when 
discharging against a 1 lb. per sq. in. back pressure. Hence, the actual 
consumption at 26 in. vacuum = 25 X 10/19 = 13.2 lb. per h.p. hr. 



324. The Function Of A Condenser Air -Pump (P, Fig. 287) 
is to produce and maintain a vacuum in the condenser by 
removing the air which enters with the exhaust steam. The 
air may leak into the exhaust system in various ways, as 
through imperfectly-packed pipe-joints and stuffing-boxes. 
Also, it may pass into the 
boilers with the feedwater, 
and thence become entrained 
with the steam-supply to the 
engine. Air if permitted to 
collect in the condenser 
would obviously prevent the 
production of an effective 
vacuum. 



2.1^0 Absolute 
12%'Pressure Difference-, 
3_ 




ILb. Or 135 Cu. Ft. Of Steam- 
l2*Jo "Pressure Difference^ 



126 Lb. Or About 2 
Cuff. Of Water 

Circulating? Pump-: 



Fig. 287. — Diagram Of Elementary Jet 
Condenser Showing Relative Volumes And 
Pressures Of Air, Water And Steam. (On 
the basis of one pound of steam.) 



Explanation. — Imagine a 
closed vessel to contain a perfect 
vacuum, and that a quantity of 
steam, unmixed with air or 
non-condensible vapors, be ad- 
mitted thereto. Then, if the steam be cooled to a temperature of 
110 deg. fahr., the absolute pressure in the condenser (Table 345) 
due to the presence of the still uncondensed water vapor would be 2.6 
in. of mercury. Or, referred to a 30-in. barometer a partial vacuum of 
30.0 — 2.6 = 27.4 in. of mercury would result. But if air had been 
mixed with the steam that was admitted to the vessel, then the air of 
itself would exert a pressure in addition to that due to the water 
vapor and would decrease the vacuum which would otherwise obtain. 
Hence, with air in the condenser, a partial vacuum of 27.4 in. of mercury 
could not result from condensation of the steam. The degree of vacuum 
actually obtainable would depend upon the quantity of air present. 



284 



STEAM POWER PLANT AUXILIARIES 



Div. 



325. The Power Required To Remove The Air And Water 
From A Condenser is, as will be shown, relatively small (Fig. 
287) compared to the power developed by the condenser: 
First estimate the power developed by 1 lb. of steam in a 
condenser under typical conditions. One pound of steam at 
an absolute pressure of 2.7 lb. per sq. in. occupies about 135 
cu. ft. The theoretical work done by the engine due to 
condensation with a vacuum of 12 lb. per sq. in., which cor- 
responds to about 2.7 lb. per sq. in. abs. then, = 135 X 144 
X 12 = 233,000 ft. lbs. for each pound of 
steam condensed. If 126 lb. (a rather-high 
value) of water are required to condense 
the 1 lb. of steam, the volume of water to 
be pumped out is, since there are 63 lb. of 
water in 1 cu. ft., 2 cu. ft. The theoretical 
work done by pump C in pumping out the 
water therefore, = 2 X 144 X 12 = 3,450 
ft. lb. for each pound of steam. If the 
volume of the air at condenser pressure is 
60 per cent, of that of the water, the work 
required to remove it, = 1.2 X 144 X 12 = 
2,070 ft. lb. for each pound of steam. The 
theoretical power required to remove 
the air and water, then, = 100 (3,450 -f- 
2,070) /233,000 = 2.4 per cent. In large 
plants the actual steam required to drive 
the condenser auxiliaries, when they are 
steam driven, amounts to about 1 to 3 per 
cent, of that required by the main engine. 

326. Gages for Measuring Condenser 
Vacuum are of two principal types: (1) 
Bourdon tube vacuum gages. (2) Mercury 




Fig. 288.— Mercury 
Vacuum Gage Which 
Reads In Inches Of 
Mercury, And In 

Pounds Per Square vacuum gages, or manometers. Both types 
Inch " usually read in inches of mercury. The 

principle of the mercury vacuum gage (Fig. 288) is similar 
to that of the barometer (See the author's Practical Heat) . 
The barometer has practically zero pressure above the mer- 
cury while with the vacuum gage the pressure to be measured 
is above the mercury. 



Sec. 327] STEAM CONDENSERS 285 

Note. — Vacuum gages of both the Bourdon-tube and the mercury 
types indicate the difference in pressure between the condenser and the 
outside air; and not the absolute pressure. Therefore the absolute pres- 
sure will be different for a given vacuum gage reading under different 
weather conditions and at different altitudes. When the barometer is 
low, a condenser will, for given cooling-water supply, efficiency and other 
conditions; give less vacuum (but the same absolute pressure) than when 
the barometric pressure is high. 

327. The Absolute Pressure In A Condenser May Be 
Computed From The Reading Of The Vacuum Gage by 

applying the following formula: 

(87) P a = Phmh 2Q [ hmv (pounds per sq. in.) 

Wherein: P a = the absolute condenser pressure, in pounds 
per square inch. P hm b = the barometer reading, in inches of 
mercury. P hmv — the vacuum-gage reading, in inches of 
mercury. 2.03 = the height, in inches, of a mercury column 
which exerts a pressure of 1 lb. per sq. in. 

Note. — If the barometer reading is corrected for temperature, the 
vacuum gage reading should be corrected for temperature also. If both 
vacuum gage and barometer use mercury columns referred to brass scales 
the error due to neglecting temperature in this formula will not be appre- 
ciable. 

Example. — A condenser vacuum-gage reads 26 in. while the barometer 
reads 29.4 in. What is the absolute condenser pressure? What is the 
degree of vacuum, as a per cent, of that which is theoretically possible? 

Solution.— By For. (87), P a = (Phmb - P hmv )/2.0S = (29.4 - 26) 4- 
2.03 = 1.67 lb. per sq. in. The degree of vacuum, as referred to that which 
is theoretically possible in this case = (26 -s- 29.4) X 100 = 88.4 per cent. 

328. The Most Profitable Average Degree Of Vacuum In 
Condenser Service is approximately as follows: (1) For 
reciprocating engines, about 88 per cent, of the barometer reading. 
This corresponds to about 26.5 in. of mercury column. (2) 
For turbines about 95 per cent, of the barometer reading. This 
corresponds to about 28.5 in. of mercury column. 

Note. — In Ordinary Reciprocating-Engine Practice it is usually 
undesirable to carry a higher vacuum than about 26.5 in. of mercury. 
There may, as hereinafter specified, be several reasons for this: (1) If 
the water discharged by the condenser is to be used for boiler-feed, its 
temperature should in many cases, for economic reasons, be higher than 
that which is due to condensation of steam in a 26.5 in. vacuum. The 



286 



STEAM POWER PLANT AUXILIARIES 



[Div. 9 



temperature due to this degree of vacuum will be somewhat less than 
120 deg. fahr. The cost of restoring heat to the 120 deg. feed-water, 
to raise it to a temperature of about 210 deg. fahr., which is suitable for 
boiler feed, may make the carrying of a high vacuum unprofitable. (2) 
But even where the supply of boiler-feed water, and the heating thereof, 
is independent of the condenser, a higher vacuum than about 26.5 in. is 
rarely justified for reciprocating engines, insamuch as the first cost of 
the installation would thereby be greatly increased. The higher the 
vacuum, the greater the cost (annual charge) of obtaining each inch 
increase of vacuum. Considerable extra expense would be incurred by 

















































= 120 




























































































-s 
















































































































































M 
















































<f 
















































-c 
















































5: 
















































c 
















































'■Ruo 






























































































s 




























/ 




















c 


























/ 






















o 


























/ 






















fc 
























/ 
















































/ 
















































/ 














































J 


























o 




















/ 


/ 














































/ 




























i 6 ! 




















/ 














































/ 






























k 5 











I 


3 








IQ 






n 




T i 




i 


D 








I 


i 



Vacuum In Inche s -Refer rec{ To 30-In. 



Absolute Back Pressures 



In Lbs. Per Sq.In. 



Fig. 289. — Graph Showing Relation Between Steam Consumption And Condenser 
Vacuum In Average Turbine Operation. (The abscissae represent percentages of 
steam consumption with 28-in. vacuum.) (Harrison Safety Boiler Works Catalog.) 



the additional precautions that would be necessary to prevent leakage 
of air through valves, stuffing-boxes and jointed connections. (3) Also, 
the initial condensation in the engine cylinder (See the Author's Steam 
Engines), due to the low temperature of the exhaust steam, would be- 
come so excessive as to more than nullify the extra advantage gained in 
lowering the back-pressure. 

In Turbine And Uniplow-Engine Practice, however, the best 
results are, aside from considerations regarding the feed-water, as noted 
above, obtained with the highest vacuum which it is possible to maintain. 
Initial condensation (as is explained in the author's Steam Turbines) 
plays no part in this case. Also, with turbines, leakage into the condenser 
can be avoided with less difficulty than in reciprocating-engine practice. 
The effect of variation in vacuum on turbine economy is shown graphic- 
ally in Fig. 289. 



Sec. 320] 



STEAM CONDENSERS 



287 



329. Table Showing Comparative Economy Of Condensing 
And Non-Condensing Operation. (Compiled by the Inter- 
national Text Book Company). 





Steam consumption per indicated 
h.p. hr. in lb. 


Per cent. 


Type of engine 


Non-condensing 


Condensing 


gained 
by con- 




Probable 
limits 


Probable 
average 


Probable 
limits 


Probable 
average 


densing 


Simple high speed. . 
Simple low speed . . 
Compound high 

speed 

Compound low 

speed. 

Triple high speed . . 
Triple low speed . . . 


40-26 
32-24 

30-22 

25-18 
24-17 
27-21 


33 
29 

26 

25- 
20 
24 


25-19 
24-18 

24-16 

20-13 
18-13 
23-14 


22 
20 

20 

18 
15 
17 


33 
31 

23 

25 
20 
29 



330. Table Showing Steam Consumption Of Condensing 
And Non-Condensing Engines. (From Gebhardt's Steam 
Power Plant Engineering.) 



Type of engine 


Pounds of steam 

per i.h.p. hr., 

non-condensing 


Pounds of 

steam per 

i.h.p. hr., 

condensing 


Per cent. 

saving due to 

condensing 

operation 


Single valve — simple . . 
Four valve — simple . . . 
Compound engine. . . . 


Average — 27 . 63 
Average — 24 . 06 
Average— 20.30 


25.7 

19.84 

12.14 


7.0 
17.5 
40.5 



Note. — The performances shown in Table 330 are for engines of a 
higher grade of construction than are those shown in Table 329. It is 
noted in Table 330 that the per cent, of saving is quite pronounced with 
the compound engine. This is due to the fact that the compound 
engines are better adapted than simple engines to handling the wide 
temperature and pressure ranges, which occur in condensing operation, 
without excessive cylinder condensation (Sec. 338) and other thermal 
losses within the engine itself. For a more complete discussion of the 



288 



STEAM POWER PLANT AUXILIARIES 



[Div. 9 



relative merits of condensing and non-condensing operation under 
different conditions see the author's Steam Engines. 

331. With Surface -Condenser Operation The Same Feed 
Water May Be Used Repeatedly (See Sec. 368). This is an 
important consideration where water must be purchased, or 
must be treated chemically to render it suitable for boiler use. 
The condensate from surface condensers may, when means 
are employed to remove the cylinder oil therefrom, be used 
over and over again as boiler feed. This will tend to prevent 
formation of scale in the boilers. Saving of fuel and reduction 
of boiler room costs will thereby result. 

332. The Advantages And Disadvantages Of Condensing 
Operation may be summarized as follows: 



Condensing 


Non-condensing 


Advantages 


Disadvantages 


Decreases engine steam consumption 
20 to 50 per cent, in large plants. 

Recovers most of the feed-water if 
surface condenser is used. 

Feed-water is available at 100 to 120 
deg. fahr. unless very high vac- 
uum is used. 

Decreases size of boiler installation. 

Increases power output of a given 
engine 25 to 95 per cent. 


Wastes most of the exhaust steam 

unless it can be used for heating. 
Must use fresh feed-water which 

may be expensive to purify. 
Requires more water-purifying 

equipment. 
Feed-water usually cold, and 

must be heated more. 
Requires larger boiler installation. 


Disadvantages 


Advantages 


Requires additional equipment, i.e., 
hot-well, condenser, cooling tower, 
pond or source of cooling water, 
vacuum pump, circulating pump, 
condensate pump, primary heater, 
etc. 

Operation is more difficult — requires 
more intelligent operators. 

No steam available for heating. 

Difficulty of keeping joints tight. 

More equipment to be kept in repair. 


Relatively low first cost. 

Operation simple — can be handled 
by less skillful operators. 

Large surplus of steam available 
for heating. 

Small steam leaks do relatively 
little harm. 



Sec. 333] 



STEAM CONDENSERS 



289 



333. Condensers May Be Classified Into Two General 
Groups: (1) Jet condensers (Figs. 290 and 291), in which con- 
densation is by direct contact. That is, the exhaust steam and 
the cooling water are mixed together. (2) Surface condensers 
(Figs. 292 and 293), in which the steam and cooling medium 
as water or air, are separated by metal walls or tubes. 
Heat is abstracted (Sec. 348) from the steam, through the 
metal, by the cooling medium. The jet condensers will be 
treated first and then the surface condensers. 



iskfckte 




•Wafer Pump 



Fig. 290. — Diagram Of Elementary Jet Condenser. 



334. There Are Three General Classes Of Jet Condensers : 

(1) Standard low-level jet condensers (Fig. 291), in which the 
water, steam and air are exhausted by pumps. (2) Siphon jet 
condensers (Fig. 294), in which the water, steam, and some- 
times air are exhausted by a barometric column. (3) Ejector 
jet condensers (Fig. 295), in which the steam and air are ex- 
hausted by the velocity or ejector effect of the cooling water. 

Note. — Jet condensers may be further classified on the basis of their 
operation as follows: (1) Parallel current condensers (Fig. 291) in which 
the condensed steam, cooling water and air flow in the same direction 
and collect at the bottom of the condensing chamber, whence they 
are evacuated by a pump, barometric tail pipe or other means. These 
condensers are used with low-vacuum installations only. (2) Counter- 
current condensers (Fig. 296), in which the condensate and cooling water 
are taken off at the bottom while the air is removed at the top. 
19 



290 



STEAM POWER PLANT AUXILIARIES 



[Div. 9 



335. The Cooling Methods Employed With Surface Con- 
densers may be classified as follows: (1) Water-cooling, in 
which heat is abstracted from the steam by circulating water- 
currents. (2) Air-cooling, in which the heat is abstracted and 
carried away by air-currents. This method is used only where 



■Seal '6 lot nol 



Cooling-Water Inlet- 




Water Seal 
for Piston Rod 



ssssssss 



Pump For Discharging* Air—-''' 
Condensate And Cooling -"Hater 




Fig. 291. — Worthington Independent Jet Condenser. 



water-cooling is impracticable, due to an insufficient supply. 
(3) Evaporation cooling, in which a cooling effect is produced 
by the evaporation of streams of water trickling on the outer 
surfaces of metal tubes through which the exhaust steam is 
made to flow. 



Sec. 335] 



STEAM CONDENSERS 



291 



■Wafer Outlet 



/Steam 




Shield to 

Prevent 

Water 

From # 

Dripping 

Into 

ToAir AirPipe 

Pump 



Fig. 292. — Elementary Double-Flow (Two-Pass) Surface Condenser. 



Condenser Shell- 



Exhaust- Circulating Water 

^faT leaves Condenser- . . 




'■Wet-Vacuum Pump Steam-Power Cylinder- Circulating Pump--' 

Fig. 293. — Typical Surface Condenser With Combined Circulating And Wet-Vacuui 
Pumps Of Reciprocating Type. 



292 



STEAM POWER PLANT AUXILIARIES 



[Div. 9 



Note. — Water-cooling is used almost exclusively in the operation of 
steam condensers. All subsequent mention of condensers in this book 
will, therefore, be limited to the water-cooled type. 



■ Cooling- Water Inlet 



-■Relief Valve 




Drain--' 



Hot 
Well- 



Overflow 



Fig. 294. — Diagrammatic Sectional View Of Buckley Siphon Condenser And 
Connections. 



335A. The Circulation Of The Water In Surface Condensers 

may be: (1) A single flow, as where the water passes (Fig. 297) 
only once through the tubing. That is, the water flows into 
the condenser at one end, through the tubes, and out at the 
other end. (2) A double flow, as where a division-plate (Fig. 



Sec. 335] 



STEAM CONDENSERS 



293 



293) is inserted so that the water passes twice through the 
tubing. That is, the water passes first through one portion 



Cooling 

Water 

■Entrance 




''•Aspirating 
Cones 



.Tail Pipe 



Outlet 
Neck 



Fig. 295. — Koerting Multi-Jet Ejector Condenser (For Large Capacities). 

of the tubes and returns through the remaining tubes, thereby 
the water travels twice the length of the tubes. (3) A multi- 
flow (Fig. 298), as where two or more division plates are in- 



294 



STEAM POWER PLANT AUXILIARIES 



[Div. 9 



Vacuum 




^Cooling Water 
\ Inlet 








Centrifugal.-' \ 
Tail Pump-''-. \ 




w/////&///syJs//&//sy/^^ ^^\sSSxv^N\\\^\^ v 



Fig. 296. — Wheeler Rectangular Rain Type Of Low-Level Jet Condenser For High 
Vacuum Service — Especially Adapted For Use With Turbines. 



Exhaust Steam ^ Discharge Outlet^ 




Drain- 



Condenser-' 



Fig. 2 9 7. — Baragwanath 
Single Flow Surface Con- 
denser. 



Cfrculafing-Wyfer Outlet . 
■Hanolholes Exhaust Inlet- 




y -Diaphragms- 

^Outlet hAirAnd Condensate Pump OhulatingrWater Inlet-' 

Fig. 298. — A Multiflow Surface Condenser. 
(This is a "four-pass" condenser.) 



Sec. 3361 STEAM CONDENSERS 295 

serted so that the water makes three or more passes through 
the tubing. 

Note. — Most surface condensers are of the double-flow or multiflow 
type. Single-flow condensers are rarely used. Double-flow or multi- 
flow condensers are used where high vacua are required, as in turbine 
operation. 

336. The Operation Of A Standard Low-Level Jet 
Condenser (Fig. 291) is* as follows: A direct-acting steam 
pump, to which the condenser is attached, removes the con- 
densate, cooling water and entrained air through a common 
cylinder, P. The cooling water enters at W. The quantity of 
water is controlled, in accordance with the quantity of ex- 
haust steam to be condensed, by the valve S, which is ad- 
justed by means of the hand-wheel H. The valve S is so 
shaped as to deliver the water in the form of a spray. The 
exhaust steam, which enters at E, mixes with the spray of 
cooling water in the chamber B. The mingled current of 
condensate and cooling water is then discharged by the pump. 

Note. — Assuming that the cooling-water supply is adequate, the 
vacuum will be maintained in a jet condenser as long as the pump keeps 
the condensing chamber free from water and air. The degree of vacuum 
in a jet condenser depends upon the water-temperature and the quantity 
of entrained air. The cooling water is, generally, drawn in by the vacuum, 
instead of being delivered by gravity. This serves to safeguard the 
engine in case of accidental stoppage of the pump. Should the pump 
suddenly stop, enough water would almost instantly accumulate to 
submerge the spray valve S. However, since insufficient water surface 
to condense the steam would then be presented, the vacuum would 
immediately be broken. Hence, flooding of the condenser and wrecking 
of the engine thereby could not occur. The engine would then exhaust 
to the atmosphere (Fig. 299) through the relief valve. 

337. To Put A Standard Jet Condenser In Service, pre- 
paratory to starting the engine to which it is attached, the 
injection valve V (Fig. 299) should first be opened. The 
pump should then be brought up to its regular running speed. 
The engine may then be started. If the pump suction is not 
sufficient to raise the water from the well the condenser must 
be primed with a small amount of cold water introduced 
through valve E. Adjustment of the spray valve S may then 
be necessary to produce the required degree of vacuum. 



296 



STEAM POWER PLANT AUXILIARIES 



[Div. 9 



Note. — If two jet condensers are to be used on the same exhaust line, 
they should be connected independently (Fig. 300, pipes D). If their 



Steam Supply Pipe- 



Engine 
Cylinder-^ 



Vacuum Gctgrz 
E 




myya-wrw .- 



Fig. 299. — Installation Of Jet Condenser With Reciprocating Engine. 

circulating systems are connected in series as by pipe A, the arrangement 
will be unsatisfactory. Condenser 1 will use water at a lower tempera- 
ture than condenser 2, and the vacuum will therefore be higher in 1. 




V"'* D 7 '*? ; : - ■'-'-"'•* f 



[Discharge Main "j 
'■Pipe L he Below Floor 



■Proper Connec tion 
Of Suction Pipe 



Fig. 300. — Showing Jet Condensers Connected In Parallel And With Injection-Water 
Connections In Series. 



If the vacuum is equalized in the two condensers by partially closing 
the exhaust steam valve to condenser 2, then nearly all the condensation 
will take place in 1. 



Sec. 338] STEAM CONDENSERS 297 

338. To Stop A Standard Jet Condenser, after closing the 
throttle of the engine to which it is attached, the injection 
valve V (Fig. 299) should first be closed. The vacuum 
breaker, D, should then be opened. The pump may then be 
stopped. 

Note. — The momentum of the flywheel will cause an engine to con- 
tinue in motion for several seconds after the throttle is closed. During 
this interval the movement of the piston will tend to produce a pumping 
effect. Hence, a slug of water may be drawn into the engine cylinder if 
the condenser pump is shut down before the injection water is shut off 
and the vacuum broken. This may occur where the engine cylinder is 
less than about 22 ft. — which is the practical maximum suction lift in 
pump operation (See Sec. 1) — above the level of the condenser, pump, or 
other source from which the water might be sucked into the engine 
cylinder. 

339. The Operation Of A Siphon Or Barometric Jet Con- 
denser (Fig. 294) is as follows: The cooling water which is 
supplied by the pump enters at E and passes downward around 
the exhaust nozzle, N, in a thin conical film. The exhaust 
steam from N is condensed within this hollow cone of falling 
water, thus creating the desired partial vacuum. The 
condensate and cooling water are discharged from the condens- 
ing chamber C by a barometric tail-pipe T. The lower end 
of the tail-pipe is submerged in a hot-well, H. In flowing 
through the neck, or constricted passage, K, the mingled 
current of condensate and cooling water acquires sufficient 
velocity to draw out the air which may be entrained with the 
steam. 

Note. — The chief purpose of this condenser arrangement is to obviate 
liability of damage to the engine by water being drawn from the tail-pipe 
into the condensing chamber and thence to the exhaust pipe. The level 
of the water in the hot-well, H, is at least 35 ft. below the condensing 
chamber. 

Atmospheric pressure cannot sustain a column of water having a height 
exceeding 34 ft. Hence it is impossible for water to get above the 
nozzle N. If the level of the injection-water supply is not more than 
about 20 ft. below the inlet, E, to the condenser, the siphoning action of 
the tail-pipe will suffice to raise the water. The pump may then be 
dispensed with after the vacuum has been formed. But if a lift of 20 ft. 
is to be exceeded the pump must be run continuously. 



298 



STEAM POWER PLANT AUXILIARIES 



[Div. 9 



340. Siphon Jet Condensers May Be Started And Operated 
Without A Pump (Fig. 301) if the injection water is to be 
lifted less than 20 ft. In such cases, the full siphoning action 
of the tail-pipe, due to gravity, is produced by a double-stage 
operation. 

Explanation. — Water is first admitted to the tail-pipe through the 
priming valve S. In falling through the tail-pipe, this current of water 

draws out the air. It thus pro- 



Mjection Wafzr Supply^ 




duces a sufficient vacuum in the 
condenser and upper part of 
the tail-pipe to draw the cooling 
water to the condenser through 
the injection pipe W. The 
priming valve, S, is then closed. 
The flow of injection water is 
regulated by means of the 
valve R. 

Note. — Sometimes, if the in- 
jection supply is to be lifted to 
a height of about 20 ft., a 
priming pipe, and valve P, lead- 
ing from the boiler-feed pump, 
must be installed. The boiler- 
feed pump may then be uti- 
lized as an aid in starting the 
condenser. 

341. The Operation Of 
An Ejector Jet Condenser 

(Fig. 295) is as follows: 
The cooling water, entering 
at W, is supplied under a 
sufficient head, due either to gravity or pump action, to send 
it through the constricted neck, K, and into the tail-pipe, 
T, at very high velocity. The exhaust steam enters at S 
and is condensed by contact with the jets of cooling water. 
The aspiratory or suction effect, due to the high velocity of the 
jets, draws the entrained air through the openings between the 
aspirating cones, A, whence it is picked up and ejected with 
the mingled current of condensate and cooling water. 

342. The Requisite Size Of Jet Condensers is generally 
determined in accordance with the records of actual experience 
in condenser operation. Numerous empirical formulas based 



Fig. 301. — 'Apparatus For Starting A Siphon 
Condenser When No Pump Is Used. 



Sec. 343] STEAM CONDENSERS 299 

on practice have been developed. A satisfactory formula, 
from Mark's Handbook, is as follows: 

(88) V = 0.00143W* + 8.25 (cubic feet) 

Wherein V = the volume of the condenser, in cubic feet. W s 
= the weight of steam, in pounds, to be condensed per hour. 

Example. — A standard jet condenser is required for a 2,000-h.p. 
engine, the steam consumption of which will be approximately 17 lb. 
per h.p. per hr. What should be the volume of the condenser? 

Solution.— By For. (88), V = 0.00143W, + 8.25 = (0.00143 X 17 X 
2,000) 4- 8.25 = 56.9 cm. ft. 

Note. — The Velocity Of The Exhaust Steam Entering A Jet 
Condenser should be approximately 600 ft. per sec. The exhaust steam 
inlet should be proportioned to produce this velocity. 

Note. — The Velocity Of The Condensate And Cooling Water 
Issuing From A Jet Condenser should be approximately as follows: 
(1) For a standard jet condenser (Fig. 291) 5 ft. per sec. (2) For a siphon 
condenser (Fig. 294) 5 to 10 ft. per sec. (3) For an ejector condenser 
(Fig. 295) 15 to 20 ft. per sec. The outlet from the bell or condensing 
chamber should be so proportioned as to produce these velocities. 

343. The Quantity Of Cooling Water Required For Jet 
Condensers depends upon the following factors: (1) The degree 
of vacuum required. (2) The heat of the exhaust steam. (3) 
The effectiveness with which the steam and water are mixed. (4) 
The quantity of air entrained with the steam. (5) The general 
efficiency of the condensing equipment. The exhaust from an 
engine generally contains considerable moisture. For practi- 
cal purposes, however, it is sufficiently accurate to assume that 
it consists entirely of dry, saturated steam. 

344. To Compute The Quantity Of Cooling Water Required 
For Jet Or Surface Condensers the following formula may be 
used: 

(89) W w = W s H ~ Tf r 32 (pounds) 

Wherein: W w = the weight of water, in pounds per hour, 
which is required to condense and cool the exhaust to a given 
discharge temperature. W 5 = the weight of steam to be con- 
densed per hour, in pounds. H = the quantity of heat, 
above 32 deg. fahr., in British thermal units, in 1 lb. of dry, 
saturated exhaust steam at the condenser pressure, as given 



300 STEAM POWER PLANT AUXILIARIES [Div. 9 

in Table 346. T/i = the temperature of the entering cooling 
water, in degrees Fahrenheit. T f2 = the temperature of the 
discharged cooling water, in degrees Fahrenheit. T fc = tem- 
perature of the condensate, in degrees Fahrenheit. (This 
temperature is the same as T f2 in jet condensers.) 

Example. — The vacuum gage of a jet condenser registers 26 in. of 
mercury. The barometer registers an atmospheric pressure of 29.4 in. 
of mercury. The cooling water enters the condenser at a temperature 
of 70 deg. fahr. The temperature of the discharge is 105 deg. fahr. The 
steam consumption of the engine is 30,000 lb. per hr. What quantity of 
cooling water is required? 

• Solution. — The absolute condenser pressure = 29.4 — 26 = 3.4 in. of 
mercury. Hence, if the barometer reading (Table 346) were 30 in. of 
mercury, the vacuum gage would show 30 — 3.4 = 26.6 in. of mercury. 
By Table 346, the total heat in the steam above 32 deg. fahr. correspond- 
ing to a vacuum of 26.6 in. of mercury = 1,112.2 B.t.u. -per lb. Hence, 
by For. (89), W„ = W S (H - T /c + S2)/(T n - T fl ) = 30,000 X (1,112.2 
- 105 + 32) -=- (105 - 70) = 890,700 lb. per hr. 

Note. — The Temperature Of The Water Discharged From A 
Jet Condenser is always lower than the temperature (Table 346) 
which is due to the condenser pressure. In high class installations it 
may be only 5 deg. fahr. below this temperature. With poorly designed 
condensers it may be 20 deg. below. But the average difference is from 
10 to 15 deg. 

345. The Operation Of A Surface Condenser is as follows: 
The cooling water is pumped through the tubes (Fig. 293) by 
the circulating pump P. The exhaust steam enters at S. The 
condensate and air are drawn out by the vacuum pump C. 
The cooling water, if admitted at the bottom, will first act upon 
that portion of the steam which is at the lowest temperature. 
This is conducive to effective transfer of heat from the steam 
to the water. It is called the counterflow principle. 

Note. — Heat Transference In Surface Condensers May Be 
Improved by preventing the condensate which forms on the upper tubes 
from falling on the lower tubes. Baffles, or rain-plates, are sometimes 
employed for this purpose. Condensers so equipped are called dry-tube 
condensers. By keeping the lower tubes comparatively dry, condensa- 
tion of the steam in the lower half of the condenser proceeds more rapidly 
than it otherwise would. Films of water enveloping the tubes serve to 
insulate them. 



Sec. 346] 



STEAM CONDENSERS 



301 





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302 



STEAM POWER PLANT AUXILIARIES 



[Div. 9 



347. The Tubes And Tube-Sheets Of Surface Condensers 

are, generally, made of such metals as are best adapted to 
resist the corrosive action of the waters which are available 
for cooling purposes. Where fresh water is used the tubes 
may be of brass, bronze, copper, aluminum-bronze, or Muntz 
metal. Where salt-water is used, tubes made of Admiralty 
metal are preferred. This is an alloy containing 70 per cent, 
of copper, 29 per cent, of zinc and 1 per cent of tin. The tube- 
sheets are generally made of brass or Muntz metal. The shell 
and fittings are commonly made of cast iron. 

Note. — The sizes of condenser tubes in common use are %-irs.., M _m -> 
and 1-in. outside diameter. The corresponding thicknesses are 20, 18 
and 16 Birmingham wire gage. Fig. 301A shows one way in which tubes 
are fastened in the tube sheets or tube head. 



Glancfs 




Tube-' 
Fig. 301A. — Worthington Standard Condenser-Tube Gland. 



348. The Quantity Of Heat Which The Cooling Water 
Will Abstract From Steam In Surface -Condenser Operation 

may be computed by the following formula : 

(90) H t = W s (# - T fe + 32) (British thermal units per hour) 

Wherein H t = the quantity of heat given up by the steam, 
in British thermal units per hour. W s = the weight of steam 
condensed, in pounds per hour. H = the total heat, above 
32 deg. fahr., in British thermal units, in 1 lb. of the exhaust 
steam at the condenser pressure, as given in Table 346. T/ c = 
the temperature, in degrees Fahrenheit, of the condensate 
leaving the condenser. 



Example. — The steam consumption of a 1,000-h.p. engine, exhausting 
into a surface-condenser, is 18 lb. per h.p. hr. The average vacuum-gage 
reading is 25.4 in. of mercury. The average atmospheric pressure, as 



Sec. 349] STEAM CONDENSERS 303 

shown by the barometer, is 30 in. of mercury. The temperature of the 
discharged condensate is 120 deg. fahr. How much heat is given up by 
the cooling-water? 

Solution. — By Table 346, the total heat, above 32 deg. fahr., in the 
steam, for a 25.4-in. vacuum with a 30-in. barometer, is 1,116.9 B.t.u. 
per lb. By For. 90, H t = W s (ff - T /e + 32) = 1,000 X 18 X (1116.9 - 
120 + 32) = 18,520,000 B.t.u. per hr. 



349. The Water-Cooling, Or Tube Surface, Required In A 
Surface Condenser may be computed by the following 
formula : 

TT 

(91) A f = — mim. (square feet) 

Wherein Af = the water-cooling, or tube surface, in square 
feet. H t = the quantity of heat to be given up by the steam, 
in British thermal units per hour, as computed by For. (90) . 
Tf 8 — the temperature of the steam, in degrees Fahrenheit, 
as given in Table 346. U = a constant from Table 350 = 
B.t.u. transferred per square foot per hour per degree tempera- 
ture difference between the water and the steam. Tji and 
T/2 = , respectively, the initial and final temperatures of the 
cooling water in degrees Fahrenheit. 

Example. — The heat to be abstracted from the exhaust steam entering 
an ordinary tj^pe of standard surface condenser, as computed by For. 
90, amounts to 18,000,000 B.t.u. per hr. The average vacuum-gage 
reading is assumed to be 25.1 in. of mercury. The average atmospheric 
pressure, as shown by the barometer, is assumed to be 30 in. of mercury. 
The cooling-water is assumed to enter at a temperature of 55 deg. fahr. 
and emerge at a temperature of 100 deg. fahr. How much tube-surface 
is required? 

Solution. — By Table 346, the temperature of the steam, for a 
25.1 in. vacuum with a 30-in. barometer, is 133 deg. fahr. By Table 350, 
the coefficient, U, of heat transference is 250. By For. (91), A/ = 
H t /{U[T fs - y 2 {T fl + T/2)]} = 18,000,000 -J- {250 X [133 - 
Y 2 (55 + 100)]} = 1,297 sq.ft. 



304 



STEAM POWER PLANT AUXILIARIES 



[Div. 9 



350. Table Of Coefficients Of Heat Transference (U, For. 
91) In Surface -Condenser Operation. 



Type of surface 
condenser 


Velocity of cooling 

water, in feet 

per second 


Value of V, in B.t.u. 

per sq. ft. per deg. 

temp. dif. between 

water and steam 


Ordinary, old style, 
standard type 


1 to 2 


250 


Modern, dry-tube type 


4 to 5 


600 



351. The Value Of The Heat Transference Coefficient, U 

(Table 350), may range from 1,000 to 3 between different 
areas of the tube-surface in the same condenser. The values 
given in Table 350 are average values. From tests made by 
Prof. Josse, of the Royal Technical School at Charlottenburg, 
it was found that the value of U is affected principally by the 
following factors: (1) The material, thickness, shape and 
cleanliness of the tubes. (2) The water-velocity through the 
tubes. (3) The steam-velocity over the tubes. (4) The quan- 
tity of condensate adhering to the tubes. 

Note. — The results of actual practice have demonstrated that sur- 
face condensers of the ordinary standard type, when attached to engines 
using 20 lb. of steam per h.p.hr., and operating with a 26-in. vacuum, re- 
quire about 2 sq. ft. of tube-surface per engine horsepower. Also, that 
dry-tube multiflow condensers, when attached to turbines using 15 lb. 
of steam per k.w.hr., and operating with a 28.5-in. vacuum, require 
about 2 sq. ft. of tube-surface per kilowatt developed by the turbine. 
Condenser practice in general indicates that from 1.25 to 2.5 sq. ft. of 
tube-surface per kilowatt are required for large modern-type installa- 
tions, while from 2 to 4 sq. ft. per kilowatt are required for the smaller 
installations of ordinary standard equipment. 

352. The Temperature "Drop" In Surface Condensers 

means the difference in temperature between the entering 
steam and the discharged cooling water. With ordinary 
standard surface condensers of the single-flow or double- 
flow type, the temperature "drop" ranges usually from 10 
to 20 deg. fahr. With high-vacuum multi-flow dry-tube 



Sec. 353] 



STEAM CONDENSERS 



305 



condensers, temperature drops of 1 to 5 deg. fahr. have been 
obtained. The temperature difference between the conden- 
sate and discharged cooling water is usually 5 to 10 deg. fahr. 
353. The Classes Of Pumps Used In Connection With A 
Condenser are: (1) Circulating pumps, or pumps used for 
forcing water through the tubes of surface condensers; or 
furnishing water to barometric or ejector-jet condensers; or 
removing water from jet condensers having dry-air pumps. 
(2) Wet vacuum pumps, or pumps used for pumping both 
condensate and air from jet or surface condensers. Wet air 
pumps for jet condensers handle the injection water also and 



z— x Steam-. A 
C_\ Turbine \jj 



Dry Vacuum 
Pump-, 




^^M^^^?^?^????^^??????^^ 



Fig. 302. — Typical Installation Of Turbine With High- Vacuum Jet Condenser And 
Pumps With 10,000 Kw. Unit. 

are sometimes called simply condenser pumps. (3) Condensate 
pumps, or pumps used with surface condensers to pump the 
condensed steam only, to a heater or receiver — usually for use 
as boiler feed. (4) Dry vacuum or air pumps, or pumps used 
for removing air only, from jet or surface condensers. (5) 
Hot-well pumps, or pumps used for pumping the hot water 
from a hot well usually to a feed-water heater. 

354. The Types Of Pumps Used As Condenser Auxiliaries 
are: (1) Direct-acting steam pumps (Fig. 293). These are used 
chiefly in reciprocating engine plants as wet vacuum pumps, 
circulating pumps or condensate pumps. (2) Rotative or 
crank-action pumps, steam or power driven (Fig. 302). These 

20 



306 



STEAM POWER PLANT AUXILIARIES 

Steam furbine-. 



[Biv. 9 




Fig. 303. — Turbine With Westinghouse-Leblanc Surface Condenser. (The equipment 
shown has been superseded by more modern designs.) 



Water-Pistons—-'. 

Pocketed Afc* 



Air 
Inlet 



■Rotating Impeller 




Hurling- _ 

Water Inlet^ £J M ! M W^ 5t ? f,0nCfr i> ■ 
f Guide Vanes' 
Compression Channel—'' 

Fig. 304. — Illustrating Principle Of Alberger Hurling-Water — Centrifugal — Air 
Pump. (As the impeller revolves, it throws streams of water out between its blades. 
Each time a stream of water passes a compression channel, a small amount or " slug " 
of water is thrown up the channel with considerable force. Air, which is admitted 
between the impeller and the channels, is caught between the slugs of water and carried 
out with them.) 



Sec. 354] 



STEAM CONDENSERS 



307 



are used chiefly for dry-vacuum pumps in either turbine or 
reciprocating engine plants. Crank-action power pumps are 
occasionally used for circulating and wet-air pumps, but steam 
drives are more common because the exhaust steam from 



■Impeller-. 



Diffusion Chambers 



Stuffing 
Box- 




Wafer Inlets ""--Discharge Outlets--" 

Fig. 305. — Alberger Hurling- Water Air Pump. 

the drives is usually needed for feed-water heating in condens- 
ing plants. (3) Centrifugal pumps (Figs. 302 and 303). 
These are the most commonly used type of circulating and 
condensate pumps in modern installations of medium and large 



Main Exhaust 



■^n 




"Pump Valves 



Water 



Fig. 306. — Parsons Vacuum Augmenter. 



capacity. (4) Hurling-water pumps (Figs. 303, 304 and 305), 
sometimes called hydro-centrifugal pumps. These are used as 
dry-vacuum pumps chiefly in turbine installations where the 
vacuum is high and the volume of air to be handled is relatively 



308 



STEAM POWER PLANT AUXILIARIES 



[Div. 9 



small. (5) Jet pumps or ejectors (Figs. 306 and 306A.). 
These are used for increasing vacuum or are built as two and 
three-stage ejectors for high- vacuum pumping service. 

Strainer Cage ■ 




Boi/er. 



■Discharge 
Opening? 



Fig. 306A. — Wheeler Two-Stage "Radojet" Air Pump Without Inter-condenser. 
(Air and water vapor are drawn from the condenser in through the suction chamber, E, 
by steam issuing from the nozzles, D, at a velocity of about 3,000 ft. per sec. The 
mixture is discharged into the diffuser, F, whence it is led to the double passage, G. 
When an intercondenser is employed, the mixture passes from F to the intercondenser 
where the steam is condensed and from which the air is led to G. Steam, delivered 
through nozzle throat, H, strikes nozzle point, J, and forms a thin sheet issuing outward 
through K and drawing air from G into the volute, L, whence the steam and air may be 
discharged into the atmosphere or into a properly-vented feed-water heater. ) 

355. The Advantages Of Centrifugal Pumps For Condenser 
Circulating Or Condensate Pumps are: (1) Low first cost. (2) 
Compactness. (3) Absence of valves and pistons. (4) High 



Sec. 355] 



STEAM CONDENSERS 



309 



Delivery 
Valves* 



Pump 
body-. 




■ Piston- 
! Drain- 
^--Entrance To Pump 



'ntrance Of 

Water And Air To 

Cylinder 



Fig. 307. — Sectional View Of Wheeler-Edwards Combined Condensate And Air Pump. 



.■Dry Air Pump 
: Connection 



Supplementary Injection Wetter 
Connection Tor Cooling Air And 
^pndensing? Vapor Mixed 
Therewith: 




Tail Pipz- X 



Fig. 308. — Condensing Chamber Of Alberger Barometric Condenser, Showing Dry Air 

Pump Connection. 



310 



STEAM POWER PLANT AUXILIARIES 



[Div. 9 



speed. This permits their being driven through direct shaft 
connection with electric motors and steam turbines. In fact 
(see Sec. Ill) centrifugal pumps are inherently high speed 
machines which renders them especially adaptable for being 
driven by motors or steam turbines which are also inherently 
high-speed machines. The same advantages apply to hurling- 
water pumps as compared to piston pumps for dry-vacuum 
pumps. 



.-Piston Rod 



M~ J ~JZ& Water 
f^-— -^1 Jacket-., 

y/^///////////////////^///^ 



■Connection 76 
Condenser 
(Air And Vapors 
PassToCylinder 
Ports Through 
Annular Passage 
Around Cylinder) 




Rotary 

Admission 

Valve 



Automatic 
Delivery Valve 



Fig. 309. — Wheeler Dry-Vacuum Pump (Single-Valve Type). (The function of the 
rotary admission valve is: (1) To connect the discharge outlet and inlet port alternately 
with opposite ends of the pump cylinder. (2) To release the compressed air in the clearance 
space at one end of the cylinder through the transfer passage to the other end of the cylinder. 
Compressed air in the clearance space is in this way released into the other end of the 
cylinder instead of back into the suction pipe where it would tend to decrease condenser 
vacuum.) 



Note. — In Order To Secure A High Vacuum With Piston Pumps 
it is essential that the clearance volume of the air-pump (Fig. 307) 
should be kept as low as possible. Also, to avoid the use of inconven- 
iently large pumps the mixture to be handled should be cooled to the 
lowest practicable temperature. Sometimes the air is re-cooled by the 
incoming condensing water (Fig. 308). In some cases a steam jet 
(Fig. 306) is used to partially compress the air before it goes to the 
pump. In others a special valve (Fig. 309) is used for reducing the pres- 
sure in front of the piston, at the end of the delivery stroke, to the 
condenser pressure. This is to obviate the loss, which would otherwise 



Sec. 355A] 



STEAM CONDENSERS 



311 



result from expansion of a portion of the compressed air down to the 
suction pressure, when the piston begins a stroke. 

Rotatory pumps (hurling or hydro-centrifugal pumps) using slugs of 
water (Fig. 304) as pistons are sometimes used where very high vacua 
are required. 

355 A. A Modern Westinghouse Turbine -Generator-Sur- 
face-Condenser Installation is shown in Fig. 309A. The 
turbine, T, is connected to a Le Blanc surface condenser, C, 
by an expansion joint, X, and a short connecting piece, J. 



Govenor* 




p. «m 



w'/^'^/j^ '- : ^'-^^':^ 



-Air Pump Discharge 
% W}f""Inlet Cooling Water Turtml 



' ^MmMiMk A - ; ^HU"~" r " 



Fig. 309A.— Westinghouse-LeBlanc Surface Condenser Installed For Service With 
Turbo-Generator Unit. 

The expansion joint is necessary to properly protect the tur- 
bine and condenser shell from excessive stress due to expansion 
when the turbine is heated by the admission of steam. The 
connecting piece is used to connect the turbine exhaust and 
the steam inlet, M, of the condenser, which may or may not 
be the same shape, and also to provide sufficient head-room 
between the turbine bedplate and the condenser shell for the 
necessary turbine supports. 

Explanation. — This (Fig. 309A) gives the most compact arrangement 
possible and requires a minimum of head-room and floor-space. The 



312 STEAM POWER PLANT AUXILIARIES [Div. 9 

condenser is placed directly beneath the turbine, T, and inside the tur- 
bine foundation. All the pumps are mounted on one shaft and driven 
by one drive. The pump unit is bolted directly beneath the shell and no 
inter-connecting piping is required. At the left is the circulating pump, 
R; in the center the air-pump, N; and at the right, the condensate pump, 
P. The pumps may be either directly driven by a motor, V, as shown or 
geared to a steam turbine. 

Ordinarily the cooling water is brought to the power house through an 
intake tunnel, /, and is discharged through a discharge tunnel, D. The 
water level should be such that the cooling water is within the possible 
suction lift for centrifugal pumps. The suction piping should be as short 
as possible to prevent air leaks and possible loss of cooling water due to the 
pump becoming air-bound. If the water levels permit, the discharge 
line, L, should be brought back to approximately the same level as the 
intake water and the end of the discharge pipe sealed by extending it, 
E, into the water. This gives a siphon system (Sec. 374) and the total 
pumping head is then only the friction head through the piping and any 
small difference in the water level, which may exist between the suction 
and discharge tunnels. 

The hurling-water air-pump, N, takes its hurling water from the 
circulating pump discharge and discharges, Q, to any convenient point 
such as the discharge tunnel. The condensate pump, P, takes the con- 
densed steam from the shell, S, and discharges it into the feed-water 
heater or feed tank. All piping — especially the circulating water piping — 
should be made as short as possible and free of sharp bends which increase 
the friction head. The circulating-water piping should have few joints 
and be free of air leaks in order to gain as much effect from siphonic action 
as possible, and thereby maintain the circulating-pump power at a 
minimum. The piping should be so arranged that no stress due to expan- 
sion will be transmitted to the pump shell. 

Note. — The Arrangement Of Condenser Pumps Shown In Fig. 
309A Is Used For The Smaller Installations where its compactness 
and simplicity make it desirable. For larger installations, separate 
pumps are used. They may be driven by one drive or separate drives 
may be provided for each pump. The air and condensate pumps 
may be combined and driven by one drive and the circulating pump by 
its individual drive. Motors or geared turbines are used also for large 
installations, the drive selected depending upon the plant lay-out. In 
some cases, the circulating pump is driven by both motor and turbine 
in order to insure added reliability and proper heat balance. (See Sec. 
212). 

356. The Principal Point To Be Observed In Caring For A 
Condenser Is To Prevent Leakage Of Air. (H. H. Kelley, 
Condensers). — Leaks may occur in cylinder-heads, valve- 
chest covers, hand-hole plates, rod stuffing boxes, flanges and 



Sec. 357] STEAM CONDENSERS 313 

screw joints in piping (both exhaust and injection pipes) and 
around valve stems. Added to these are the piston rod and 
valve-stem stuffing boxes on the engine and any bonnets that 
may lie below the exhaust valves. The leaks are not readily 
detected because the pressure is on the outside and the air is 
consequently trying to get in. A lighted candle or match 
held close to the joint where the leak is suspected is about the 
simplest method of locating them. The suction and discharge 
valves of condenser pumps should be examined regularly as 
there is no direct way of detecting loss of vacuum due to poor 
valve action. 

357. Strainers Should Be Placed At The Ends Of The 
Circulating Water Suction Pipes whether the condenser takes 
water directly from a creek or pond or from an intermediate 
reservoir. Openings in the spraying device of a jet-condenser 
are sometimes comparatively small and these may become 
clogged with bits of foreign matter that might readily pass 
through the suction valves. The tubes of surface condenses 
may also become clogged by foreign matter and thus decrease 
the flow of circulating water. When a jet condenser fails to 
get sufficient water, first examine the strainer, then the spray, 
which can usually be reached by removing the small manhole 
plate on the condenser chamber. If trouble is not found at 
these points, examine the pump ports, the suction valves, dis- 
charge valves and, lastly, the piston or plunger. If no obstruc- 
tion is found, the difficulty will be due either to leakage of air 
or the heating of the condenser caused by receiving more steam 
than it is capable of condensing; in other words, the condenser 
is too small. 

358. Should The Condenser Vacuum Suddenly Decrease 
While Running, it will probably be due to an increased load 
on the engine and the correspondingly greater volume of 
steam entering the condenser. The amount of injection 
water which was formerly sufficient would then be too small 
for the weight of the steam which is to be condensed. The 
obvious remedy is to open the injection valve. If this does 
not restore the vacuum, slowly increase the speed of the pump, 
always watching the vacuum gage, while making these adjust- 
ments. If the loss in vacuum is due merely to a larger amount 



314 STEAM POWER PLANT AUXILIARIES [Div. 9 

of steam, these adjustments will restore it. If the vacuum 
decreases slowly, a little each hour of the day, it indicates 
leakage of air, a leaky piston and valves or stoppage of the 
water passages somewhere between the suction strainer and 
the discharge valves. The several joints and the stuffing boxes 
may be examined for air-leaks in from 5 to 10 min. while 
running. But an examination of the valves, spray and pump 
cylinder can only be made after shutting down the condenser. 

Note. — If The Condenser Has Become Hot It Will Not Work 
Until It Is Cooled. As it is necessary to bring steam in contact with 
a colder body in order to condense it, should the temperature in the 
condenser rise nearly to that of atmospheric exhaust steam, condensation 
will take place slowly and the vacuum can be re-attained only gradually 
as the condenser cools again. 

359. When The Atmospheric Relief Valve Of A Jet Con- 
denser Is Open And The Engine Is Running Non-Condensing, 
Proceed As Follows To Restore The Vacuum And Condensing 
Operation. — After locating and removing any cause of diffi- 
culty, the pump may be started and the injection valve opened. 
The temperature will thus be lowered to that of the condensing 
water. With an assistant at the atmospheric relief valve, 
speed up the pump and give the condenser more water. Then 
slowly open the stop valve in the exhaust pipe, having the 
assistant close the relief valve at the same rate as that at which 
the stop valve is opened. When the relief valve is nearly 
closed, it will close itself due to the vacuum which will then 
have been produced in the exhaust pipe, and the engine will 
run condensing again. The injection valve may then be 
partly closed and the speed of the pump reduced a little, always 
keeping watch of the vacuum gage while making these ad- 
justments. The object is to use as little water as possible 
and run the pump as slow as possible and still maintain the 
desired vacuum. 

Note. — The Above Suggestions Apply Also To Surface Condens- 
ers. The only difference is that in some surface condenser plants, the 
air pump and circulating pump are regulated separately. Increasing 
the speed of the air pump is equivalent to increasing the speed of the 
pump in the jet condenser. Increasing the speed of the circulating pump 
has the same effect as opening the injection valve in the jet condenser. 



Sec. 360] STEAM CONDENSERS 315 

360. It Sometimes Happens That The Vacuum Is Consider- 
ably Below That Which Corresponds To The Condenser 
Temperature, i.e., the temperature of the condenser may 
correspond to a vacuum of 26.5 in. while the highest vacuum 
which can be maintained is 25 in. In most instances this 
will be due to air in the condenser and a thorough search for 
leaks should be made, provided the vacuum gages and ther- 
mometers are known to be correct. It is practically impossible 
to maintain a condenser system sufficiently free from air 
that the vacuum-gage reading will correspond exactly to the 
temperature. A reasonable or allowable difference between 
the vacuum gage reading and the vacuum corresponding to 
the condenser temperature, as found in a steam table, is 
about 0.5 in. mercury column. 

361. The Adjustments And Care Of The Barometric And 
Ejector Jet Condensers consist largely of regulating the in- 
jection valve and preventing leaks. When a dry vacuum 
pump is employed in connection with a barometric condenser, 
it may need repair or the speed may require changing in case 
of difficulty in maintaining the vacuum. Ordinarily these 
pumps are provided with governors, the speed being changed 
quickly, when need be, by adjusting the governor. 

362. With Surface Condensers, Leaky Tube Ends And 
Fouling Of The Tubes Both Inside And Out May Give Trouble. 
This condition shows itself in a gradually falling vacuum. 
Increase of the speed of the air and circulating pumps affords 
but temporary relief. The remedy is in thorough cleaning. 
The inside of the condenser may usually be cleaned with a 
hose and ordinary city water pressure. A nozzle of pipe 
small enough to go inside the condenser tubes is fitted to a 
hose. A thick leather washer around the nozzle may be used 
to prevent the water from squirting back and wetting the 
operator when the nozzle is inserted in the tubes. If a valve 
is placed near the nozzle, the work may be done by one man. 
After removing the head of the condenser, the nozzle is 
pushed in and the water is turned on. If the water fails to 
clean out the tubes, a rod having a spiral end like an auger 
may be used to scrape the tubes clear after which they may 
be rinsed with water as described above. 



316 STEAM POWER PLANT AUXILIARIES [Div. 9 

363. When Grease Accumulates On The Outside Of The 
Condenser Tubes it may be removed by boiling the condenser 
out with lye: Remove the handhole plate and put in several 
cans of lye, 6 or 8 lb. for a 500 h.p. condenser and 12 to 15 lb. 
for a 1,200 to 2,000 h.p. condenser. Provide a small live steam 
pipe reaching well down into the condenser. Fill the condenser 
with water. Heat the water to the boiling point with the 
steam pipe and permit it to stand for 18 to 24 hr. The grease 
will then run out with the water — mostly in the form of soap. 

364. An Index As To The Condition Of Joints And Stuffing 
Boxes Of Any Condenser can be obtained by noting the loss 
of vacuum after shutting down. If all the connections, stuff- 
ing boxes, and joints are reasonably tight, the loss of vacuum 
should not exceed 2 in. per hr. 

365. The Following Material On Condenser Selection And 
Economics is based largely on an article, Application of 
Condensers, by F. A. Burg which appeared in The Electric 
Journal for Dec, 1920. 

366. Features Which Should Be Considered When Select- 
ing The Type Of Condenser To Use for a given installation are 
these: (1) The space available. (2) The boiler feed problem. 
(3) The cooling water. (4) Maintenance. (5) First cost. In 
most cases, by a general survey of these items, the selection 
can be made without resorting to refinements and calculations. 
If, however, such a survey shows that there is little choice 
between types, then each type of condenser should be con- 
sidered individually. The most economical size of each type 
should be determined, and then these should be compared 
rather than arbitrarily selected. The recommended general 
procedure in making a selection is to determine for each 
condenser type under consideration the excess operating and 
installing costs involved. Then when these have been ascer- 
tained the propositions should be summarized and balanced 
against one another. The excess total annual operating and 
maintenance costs should be capitalized at a reasonable percent- 
age and the resulting amount added to the first cost of the 
condenser that has the excess operating cost. This total is the 
amount that it is justifiable to pay in initial cost for the 
condenser which effects the saving. 



Sec. 367] 



STEAM CONDENSERS 



317 



Example. — Condenser A costs SI, 000 and its total annual operating 
(power and maintenance) cost is $400. Condenser B costs $700 and its 
total annual operating cost is $500. Which of these condensers is the 
more economical? 

Solution. — Difference in operating (power and maintenance) cost = 
$500 — $400 = $100 annually. Assume a total annual fixed charge 
(rental cost of space occupied, interest, depreciation, taxes and insurance) 
of 15 per cent, on the investment. This $100 annual saving corresponds 
to a saving in investment of $100 -4-0.15 = $666.70. Therefore it is 
economical to pay $666.70 more for condenser A than for condenser B. 
But A cost only $300 more than B. Hence A is the best investment. 
Another method of arriving at the same conclusion is to tabulate the 
data thus: 



Item 


A 

First cost = 
$1,000 


B 

First cost = 
$700 


Operating cost 

Fixed charge @ 15 per cent 


$400 
150 


$500 
105 


Total annual charge 


$550 


$605 



Thus the data shows that the yearly or annual cost of A is $605 - $550 = 
$55 less than that of B. This $55 annual-cost saving would justify 
an increase in investment of $55 -J- 0.15 = $366.70. That is: $366.70 + 
300 = $666.70. 

367. The Amount Of Floor Space And The Head Room 
Available Are Rarely Deciding Factors In Selecting Con- 
densers. — Surface condensers require more floor space than 
do jet condensers, especially when allowance is made for the 
space required for removing the tubes. In a new plant, space 
for a surface condenser can be provided without difficulty, but 
frequently turbine foundations must be specially designed 
to accommodate the condenser. The head room required 
for either low level jet or surface condensers is about the 
same, if the possible variations in design, such as different shell 
proportions or the use of twin units, are recognized. Generally 
the question of space is not of primary importance. However, 
the difference in the cost of the installation due to the differ- 
ence in space occupied, if any exists, should be reflected in 
the cost analysis of the problem. 



318 STEAM POWER PLANT AUXILIARIES [Div. 9 

368. The Quality Of The Available Feed Water Is Often 
An Important Factor In Condenser Selection. — The surface 
condenser recovers the distilled condensate for boiler feed 
while the jet does not. There are relatively few natural 
waters which do not contain sufficient solid matter, either in 
suspension or solution, to form scale in boilers. Some waters 
contain minerals that form a hard scale. Others, with just as 
high a mineral content, form a soft easily-removable scale. 
The questions of treating feed water, what minerals are most 
objectionable and methods of cleaning boilers cannot be 
discussed here, but many feed waters have to be treated. 
The methods of obtaining good feed water vary from a chemical 
treatment of all of the feed water to the recovery of the condensate 
with a surface condenser, and treating only the make-up water. 

Note. — Although Surface Condensers Should Deliver Pure 
Distilled Water To The Feed Heater, They Often Do Not Do So. 
The purity of the water depends on the tightness of the tube packing and 
the condition of the tubes themselves. If the tubes leak the feed water 
will be adulterated by the amount of the leakage. Hence, frequent 
electrical or chemical tests of the condensate should be made to determine 
its quality. 

369. The Character, Quantity And Source Of The Cooling 
Water Are Important Factors In Condenser Selection. — A con- 
densing plant requires for condensing water alone from 25 to 
100 lb. of water per lb. of steam condensed. A plentiful supply 
of water at a low temperature, and at such elevation as to 
involve minimum pumping power expense, is desirable. Natu- 
ral heads are desirable but not often available for steam plants. 
Where the water supply is limited, an artificial cooling system 
can be installed (see Div. 10). The amount of water then 
circulated will depend on the cooling range that can be effected 
by the cooling system and not on the type of condenser 
employed. 

370. Cooling Towers And Spray Ponds (see also Div. 10) are 
both used for artificial cooling. The rise in the temperature 
of the cooling water must be kept within the cooling range of 
the tower or pond, since the water has to be cooled in the tower 
or pond by the amount that it has been heated in passing 
through the condenser. For the average conditions of tempera- 



Sec. 371] STEAM CONDENSERS 319 

ture and humidity, say 70 deg. fahr. air temperature and 70 
per cent, humidity, the cooling range for a natural-draft tower 
or a spray pond, single-spraying, is usually assumed to be from 
14 to 16 deg. fahr. and, for a forced-draft tower or a pond with 
double-spraying system, from 22 to 25 deg. fahr. This means 
that the ratio of water to steam would be between 60 and 70 
to 1 in the first case and about 40 to 1 in the latter. 

371. With Surface Condensers Probably Not More Than 
90 To 93 Per Cent. Of The Boiler Feed Will Be Returned To 
The Boilers. The Rest Will Have To Be Made Up.— This 
make-up water will, with surface condensers, have to be treated. 
But the expense of such treating is small as compared to the 
expense of treating incurred with jet condensers, where all the 
feed must be treated. There will also be a loss of heat in the 
feed when jet condensers are used even if the feed is taken from 
the discharge of the condenser because the temperature of the 
condensate from a surface condenser is higher than the tem- 
perature of the discharged cooling water from a jet condenser. 

372. When Investigating The Feed-Water Phase Of The 
Problem it will therefore be necessary to find out the excess 
cost of treating the feed, the amount chargeable to the jet 
condenser for the loss of heat in the feed water and the excess 
cost of the treating plant. The cost of treating is variable. It 
depends on the nature of the water to be treated. Ordinarily 
the cost does not exceed fifteen cents per thousand gallons. 
The loss of heat involved can be reduced to the amount of 
steam required to raise the temperature of the feed water to 
that of the condensate in a surface condenser. After this has 
been determined the cost of generating this steam may be 
ascertained. The cost of a treating plant will depend on the 
method used and the amount of water to be treated. With all 
these items known another step in the analysis has been 
completed. 

373. The Effects Of Bad Water on jet condensers are of less 
moment than on surface condensers. In jet condensers the 
parts subject to corrosion can be replaced more cheaply. The 
tubes in a surface condenser will last indefinitely, if the water 
is noncorrosive. But, surface condensers are frequently used 
where only corrosive water is available. When the water is 



320 STEAM POWER PLANT AUXILIARIES [Div. 9 

quite bad, the tubes must be made of a special metal and even 
then may last only a short time. When the water is thus bad, 
although it may be highly desirable to save the condensate, 
the cost of doing this may not compare favorably with the cost 
of boiler feed from some other source. 

374. The Most Important Phase Of The Cooling Water 
Problem Is The Cost Of Handling The Water under the condi- 
tions that may exist in the power plant. The jet condenser, 
by reason of its ability to realize a lower terminal difference 
(difference between the temperature of the exhaust steam and 
that of the outgoing cooling water) does not require as much 
water under average conditions as does the surface condenser. 
This, however, does not mean that it will require less power. 
With the jet condenser its circulating pump has to pump all the 
water out of a partial vacuum which corresponds to about 30 
ft. head. In addition it must discharge against an external 
discharge head that is never less than the discharge head on 
the surface condenser. The external head consists of the 
static lift plus the friction. This means that the jet condenser 
always has a pumping head in excess of thirty feet, whereas the 
surface condenser may not require a head greater than that 
due to condenser and pipe friction. The head would not be 
greater than that due to condenser and pipe friction where the 
cooling water is taken from a body of water and discharged 
back at the same level provided that the whole system is so 
sealed that the full siphonic effect is realized. Such installa- 
tions occur frequently. See Figs. 303 and 309A. 

375. When The Discharge Level Is Higher Than The 
Circulating Pump (Figs. 310 and 311), which condition is 
ordinarily encountered in spray-pond installations, the advan- 
tage of lower pumping head is also with the surface condenser 
because the surface condenser can under this arrangement take 
advantage of the balanced leg in the circulating system while 
the jet condenser cannot. 

Example. — Assume a cooling-tower installation with the level of the 
cold well ten feet above the circulating pump. There will be a 10-ft. 
positive head on the pump for the surface condenser (Fig. 310). This 
10 ft. can be credited because, under static conditions, the level 
of the water in the discharge pipe would be 10 ft. above the pump. 



Sec. 3761 



STEAM CONDENSERS 



321 



But a jet condenser (Fig. 311) cannot take advantage of this head because 
it would have to pump the water against a 10-ft. head in addition to the 
internal head due to the vacuum. From this it is evident that, in most 
cases, the circulating pump of a surface condenser pumps against a lower 
head than does the pump of an equivalent jet condenser. 



Surface 
Condenser-, 




vAAa^aAaa 

A/\/ySA,A/vy\ r 
wMmAa./ 
a, A/y\ AA, /yVv 
sAAAaaaaa^ 
a a /\ a a, aa a a, 
AA.A,A,AA,AA/ 



w?mmm 






,.</'rcu/cff/hgr 
Pump 



Fig. 310. — Showing Pumping Head Of 
Surface-Condenser Circulating Pump. 



tf= 






vaaaaaaaa, 
aaAaaaAa.A 

\A,AMAA A A 
lAA AAA A A A A 
A A.A A AA,A A A. 
A.AAAA.AAA/' 
VAA.AAAAAA 




Fig. 311. — Showing Pumping Head Of' 
Jet-Condenser Circulating Pump. 



376. With The Jet Condenser, The Ratio Of Water To 
Steam Is Fixed For A Given Vacuum Whereas With The 
Surface Condenser This Ratio May Be Varied To Suit The 
Conditions. — The vacuum obtainable depends, with a surface 
condenser, on a variety of factors, including the rate of heat 
transfer through the tubes, the velocity of the circulating 
water, and the size of the condenser. Hence it is possible to 
select a number of surface condensers, each with a different 
ratio of water circulated to cooling surface, that will produce 
approximately the same vacuum with a given amount of 
steam. But with a jet condenser, the quantity of water 
required is practically fixed when the quantity of steam and 
the vacuum are specified. This is due to the fact that most 
jet condensers realize a terminal difference in temperature of 
between 5 and 8 deg. fahr. That is, the rise in the tempera- 
ture of the circulating water and the ratio of water to steam 
will be practically the same for all jet condensers producing a 
given result. On the other hand, with the surface condenser, 
the ratio of water to steam may be varied to suit the conditions 
of different pumping heads and the necessity for conservation 
of auxiliary power. 

21 



322 STEAM POWER PLANT AUXILIARIES [Div. 9 

377. It Is Generally Recognized That For High Circulating- 
Water Heads There Should, To Insure Minimum Surface - 
Condenser Operating Expense, Be A Lower Ratio Of Water To 
Surface Than For Low Heads. — Thus, where the circulating 
water must be pumped against a high head, it is economical to 
decrease the amount of water to be pumped by installing a 
larger surface condenser. That is, the auxiliary power and 
therefore the operating expense may be decreased by incurring 
a greater initial expense for a larger condenser. No such 
economic adjustment is possible with a jet condenser. Since, 
however, for a given service, the jet condenser usually uses 
less water than the surface condenser, it may approach the 
surface condenser in auxiliary economy when the head is high. 
The handicap of the jet condenser circulating pump of having 
to pump against a greater head will then, where the head is 
high, be offset by the lesser amount of water to be pumped. 

378. A Comparison Of The Power Requirements Of The Jet 
Vs. The Surface Condenser Should Not Be Based Solely 
On A Consideration Of The Quantities Of Water Circulated 
And The Heads Existing. — There should be considered also: 
the facts that the jet condenser discharge pump is inherently 
less efficient than a pump not discharging from a vacuum, and 
that the jet condenser must have a larger air pump than the 
surface condenser. Against all these jet-condenser handicaps 
of less efficient pumps greater heads and more power for the 
air pumps, the jet condenser has the advantage of less water 
to circulate. However, for most installations, the jet con- 
denser requires more power for drive. The amount of this 
excess depends on the discharge head, the type and capacity 
of the air pump, and the vacuum at which the condensers are 
compared. With this excess determined, an excess charge 
in operating expense can be made. This should be taken at a 
fair rate per horse-power-hour for the total number of hours 
per year the condenser will be in service. This data provides 
another item for the final comparison. 

379. The Type Of Drive For Condenser Pumps, Whether 
Electric or Steam, depends entirely on the use that can be 
made of the exhaust steam. If other steam-driven auxili- 
aries, such as drives for stokers, fans and boiler-feed pumps, 



Sec. 3801 STEAM CONDENSERS 323 

furnish sufficient exhaust steam to heat the feed water (Sec. 
265) it will not be necessary nor economical to have the con- 
denser auxiliaries steam driven. In accounting for the excess 
steam required by steam drives, it is customary to disregard 
a charge if all the steam can be used advantagously in heating. 
If the condenser auxiliaries are motor driven, the charge is 
usually determined by taking the water rate on the turbine 
from which the motor derives its power, allowing for all 
electrical losses, and thus arriving at the equivalent steam 
consumption per horse-power of the motor load. In most 
plants this will run from fifteen to twenty pounds of steam 
per horse-power-hour. The charge for the excess steam can 
then be determined from the cost of producing the excess 
steam that is required. 

380. In First Cost The Jet Condenser Has A Decided Advan- 
tage Over The Surface Condenser. — A jet condenser usually 
costs about half as much as an equivalent surface condenser. 
A standard low-level jet condenser for a 10,000 k.w. turbine will, 
at the present prices, cost about $30,000 delivered and erected. 
A surface condenser for the same turbine will cost about 
$65,000. It is this great difference in first cost that often 
renders the installation of the jet condenser justifiable. Such 
a wide difference in first cost will offset a considerable amount 
of capitalized savings. 

381. The Cost Of Maintaining Pumps Of A Jet Condenser 
Will Not Be As High As For The Surface Condenser.— This 
is because the jet condenser has only two pumps and the 
surface has three. But the difference in these repair costs is so 
slight that it can usually be neglected. 

Note. — Pump Runners May Last From A Few Months To Several 
Years, depending on the kind of water being pumped. Hence it is 
infeasible to quote any general data on the cost of making runner 
replacements. 

382. As To The Relative Costs Of Cleaning Jet And Surface 
Condensers: the jet requires practically no attention, often 
operating for years without being opened. But a surface 
condenser must be cleaned frequently to prevent a serious loss 
of vacuum. The loss of vacuum is due to the decrease of the 



324 



STEAM POWER PLANT AUXILIARIES 



[Div. 9 



rate of heat transfer caused by dirty tubes and a consequent 
restriction in the flow of water through the condenser. The 
frequency of cleaning and its cost vary with quality of the 
cooling water. In some plants condensers must be cleaned 
weekly. In others they are cleaned monthly. Often they 




"Handle 



Fig. 311 A. — Worthington Hydraulic Tube-Cleaner. 

are not cleaned at regular intervals but only after a definite 
loss in vacuum has been observed. The cost of cleaning 
varies with: (1) the character of the deposit on the tube, (2) the 
method of cleaning, (3) facilities for handling the water box 
covers, (4) the price of labor. This cost may vary from 2^ 
or 3^ per sq. ft. per yr. to 15^ or 20^ per sq. ft. per yr. 




"Removable 
Cap 



Fig. 31 IS. — Section Through Ball-Joint Of Worthington Hydraulic Tube-Cleaner. 



Note. — Built-In-Tube-Cleaning Equipment For Surface Con- 
densers (Fig. 311 A) is very useful and economical where condensers use 
water which is economical where condensers use water which is apt to 
leave a deposit of vegetable matter, mud, or slime. In the device shown 
in Fig. 311 A, a ball nozzle, B, may be attached to each manhole plate of 
the condenser. Through it a cleaning nozzle, A r , may be inserted. 
Water is led to the nozzle through the hose, H, at a pressure of 250 lb. 
per sq. in. The nozzle can be swung to different positions by handle 



Sec. 383] STEAM CONDENSERS 325 

A on the outside of the condenser. The nozzle delivers about 70 gal. per 
min. which, it is claimed, will remove the mud from a completely-filled 
tube. With this device, the condenser tubes can be cleaned in a 
relatively-short time. 

383. An Important Item Of Surface -Condenser Expense, 
Is The Replacement Of Tubes. — Tubes may last several years 
or they may last only a few months, depending on the compo- 
sition of the tubes and the character of the water. Instances 
have been recorded where tubes have lasted for fifteen years 
but this is exceptional. In industrial communities, where 
the water is liable to be contaminated with sewage and refuse, 
five or six years is often an average life. Here again the limits 
are so wide that general figures would be misleading. At 
the present price of tubes assuming a 5-yr. life it would cost 
about 18c 7 per sq. ft. per yr. for tubes alone, disregarding the 
cost of extra ferrules and packings, and the cost of installation. 
These latter items would increase the above-quoted value to 
at least 30c per sq. ft. per yr. for replacements. 

384. The Selection Of The Proper Condenser To Serve 
A Steam-Driven Prime Mover Is A Problem In Power-Plant 
Economics. — A condenser, regardless of type, is installed in a 
modern power plant only because of the reduction which it 
effects in the cost of power. Operating condensing, as com- 
pared with non-condensing operation, cuts the cost of power 
approximately in hah in large turbine stations. It is therefore 
essential that great care be exercised in making the selection, 
so that the full saving from condensing operation may be 
realized. Thus the problem of condenser application reduces 
itself to. a calculation to determine which condenser equipment 
will produce power at the least cost. The following illustrative 
example explains the method : 

Example. — Decide between a surface and a low-level jet condenser for 
a 10,000 kw. plant, under the following conditions: 200 lb. per sq. in. 
steam pressure; 100 deg. fahr. super-heat; space no consideration; boiler 
feed treating costs 12£ per 1,000 gal., including chemicals and attendance; 
abundant cooling water, available at a head of 10 ft., external to the 
condenser; surface condenser maintenance including cleaning and 
replacing tubes, 35^ per sq. ft. per yr. ; maintenance on pumps, the same in 
either case; plant operating 7,000 hr. per yr. at an average condensing 
load of 100,000 lb. per hr. ; condensers to be selected on the basis of 75 deg. 



326 



STEAM POWER PLANT AUXILIARIES 



[Div. 9 



fahr. water and 28.25 in. vacuum; condenser pumps to be motor driven 
because no exhaust steam is required for heating the feed water. 

Solution. — First, select the sizes of the condensers: The installation 
will require a jet condenser circulating 12,000 gal. per min. or a surface 
condenser having 15,000 sq. ft. of surface and circulating 17,500 gal. per 
min. to give the required performance. The ratio of water to surface 
for the surface condenser has been taken in accordance with common 
practice for this low-head condition. The jet condenser will require 
340 h.p. for its drive and the surface condenser will require 230 h.p. 
The water rate on the main turbine is about 101b. perb.hp.h.r. Assuming 
that the motors will receive their power from the main unit at an overall 
transmission efficiency of 82 per cent., including generator, motor, trans- 
former and line losses, the steam per h.p. hr. chargeable against the pumps 
will be: 10 -s- 0.82 or 12.2 Z6. per h.p. hr. It has been assumed that the cost 
to generate the steam will be 35 j£ per 1,0001b. and the motor-driven pumps 
are charged on this basis. 

For this plant it is assumed, if a jet condenser is used, that the 
water for boiler feed would be taken from the discharge side of the 
condenser thus realizing the advantage of the higher temperature. 
When condensing 100,000 lb. of steam per hr. the condenser chosen 
will have a discharge temperature of 91-deg. fahr. and the hotwell tempera- 
ture of the surface condenser would be about 92 deg. fahr. There is 
such a slight difference in these temperatures that the heat lost in the 
feed when using a jet condenser as compared that with the surface 
condenser may be disregarded without serious error. 

The costs of operating the surface as against the jet condenser may be 
summarized thus: 



Items 


Surface 


Jet 


Cost of boiler feed per year 

Cost of power for pumps 

Maintenance, surface. . . 

Pump maintenance, same 


$ 785 
6,875 
5,250 


$ 9,800 
10,160 


Totals 


$12,910 
$ 7,050 


$19,960 




Saving in favor of surface 




Capitalized against jet @ 15% . . . . 
First cost of condensers 


65,000 
2,000 


47,000 
30,000 
15,000 




Cost of water treating plant 




Totals 


$67,000 


$92,000 









From the above tabulation it is evident that the surface condenser has the 
advantage over the jet. Based on the saving effected by the surface con- 
denser, we could afford to pay $92,000 for it with the water treating 



Sec. 384] STEAM CONDENSERS 327 

equipment, whereas it costs only $67, 000. This same conclusion could 
also have been reached by calculating net savings instead of capitalized 
savings. But the former method is usually preferable since it indicates 
directly whether the prices asked for the equipment are justifiable. 
The condensers used in this comparison were selected on the basis of 
common practice. 

QUESTIONS ON DIVISION 9 

1. How does a condenser save steam? Increase power output? 

2. Explain the operation of Newcomen's condensation engine. How did Watt 
improve this engine? 

3. What is the function of a condenser air-pump? Why is it necessary? 

4. How does the power required by the condenser auxiliaries compare with that 
developed by the condenser? 

5. How is condenser vacuum measured? How is it affected by weather conditions? 

6. What is approximately, the most profitable vacuum for reciprocating engines? 
For turbines? Why the difference? 

7. Give a few advantages and disadvantages of condensing operation. 

8. How may condensers be classified? Name three classes of jet condensers. 

9. What is the cooling medium commonly employed in condensers? 

10. Name three classes of surface condensers. 

11. Explain the operation of a standard low-level jet condenser making a sketch of 
the main parts. How is water in this condenser prevented from getting into the engine? 

12. How is a standard jet condenser started and stopped? 

13. How should two jet condensers be connected when they are used on the same 
exhaust line? 

14. How is the air removed in the ejector jet condenser? 

15. What are the functions of the tail pipe of a barometric condenser? Explain 
how and under what conditions a siphon or barometric condenser may be started and 
operated without a pump. 

16. What, approximately, should be the velocity of the entering steam in a jet con- 
denser? Velocity of cooling water issuing from a standard jet condenser? From an 
ejector jet condenser? From a siphon condenser? 

17. On what does the quantity of cooling water for a jet condenser depend? 

18. Explain by a sketch the operation cf a double-flow, dry-tube, horizontal, surface 
condenser having separate air and condensate pumps. Explain the counterflow princi- 
ple as used in this type of condenser. 

19. What is the composition of the tubes, tube sheets and shells of most surface 
condensers which use salt water for cooling? 

20. What factors determine the heat-transfer coefficient of a surface condenser? 
Give approximate values for the tube surface required per turbine kilowatt developed. 

21. What is meant by temperature "drop" in a condenser? Give representative 
values for it. 

22. What kinds of pumps may be used as condenser circulating, condensate and dry- 
air pumps? 

23. How are reciprocating pumps designed for high-vacuum pumping service? 

24. Where may leaks occur so as to impair condenser vacuum? How may they be 
located? 

25. State the usual causes of the failure of a jet condenser to get sufficient water and 
explain remedies. 

26. How may an operator know whether decreased vacuum is due to leaks or to 
increased load? 

27. How may a jet condenser be started if it has become hot? 

28. How is a steam table used in determining whether or not a condenser is reasonably 
tight and efficient? What is another way of testing for tightness? 

29. Explain methods of cleaning condenser tubes inside and outside. 



328 STEAM POWER PLANT AUXILIARIES [Div. 9 

30. What are the relative importance of space, head-room and maintenance changes 
in selecting a condenser? 

31. How may cooling-water supply affect the selection of a condenser? 

32. What is a typical value for the cost of purifying feed water? How does this 
value enter into condenser selection? What per cent, of the original boiler feed is 
ordinarily recovered by a surface condenser? 

33. Why must a jet-condenser circulating pump always work against a 30 ft. greater 
head than a surface-condenser circulating pump under the same conditions? How 
may the pumping economy of a jet condenser equal that of a surface condenser in spite 
of this fact? 

34. How do the first costs of surface and jet condensers compare? Cost of cleaning? 

35. Explain with an example how the economic advantage of two condensers may be 
compared on the basis of capitalized saving. 

PROBLEMS ON DIVISION 9 

1. Steam is admitted to a steam turbine at 450 deg. fahr. It is exhausted at 225 deg. 
fahr. when running non-condensing and at 80 deg. fahr. when running condensing. 
What are the greatest possible thermal efficiencies when running condensing and non- 
condensing? 

2. If an engine has a mean effective pressure of 78 lb. per sq. in. running non-con- 
densing, what will be the saving in power due to condensing operation with a 26.5 in. 
vacuum? 

3. A turbine consumes 22 lb. of steam per h.p.hr. at 185 lb. per sq. in. abs. when run- 
ning non-condensing, exhausting against 1 lb. per sq. in. back pressure. What will be 
its steam consumption if its exhaust is condensed in a 29 in. vacuum? 

4. The vacuum gage of a condenser indicates 27 in. of mercury. The barometer 
registers 29.8 in. of mercury. What is the absolute condenser pressure in inches of 
mercury? In lb. per sq. in.? What per cent, of the vacuum possible at the prevailing 
barometric pressure does this represent? 

5. A siphon jet-condenser is required to condense 10,000 lb. of exhaust steam per 
hour with 36 lb. of cooling water per pound of steam. The velocity of the discharge 
through the tail-pipe is 5 ft. per sec. What should be the volume of the condenser? 
What should be the diameter, in inches, of the tail-pipe? 

6. The vacuum gage of a jet condenser registers 27 in. of mercury. The barometer 
registers 30 in. The temperature of the cooling water at entrance is 80 deg. fahr. The 
temperature of the discharge is 105 deg. fahr. The engine exhausts 10,000 lb. of steam 
per hour. What is the temperature-difference between the discharge-water and the 
entering steam? How many gallons of cooling water are required per minute? 

7. The vacuum gage of a surface condenser registers 28 in. of mercury. The barome- 
ter registers 29.5 in. The condenser receives 10,000 lb. of exhaust steam per hour. 
The cooling water enters at a temperature of 67 deg. fahr. and leaves at a temperature 
pf 87 deg. fahr. The temperature of the condensate is 85 deg. fahr. How much cooling 
water is used per hour? 

8. A surface condenser condenses 150,000 lb. of steam per hr. at an absolute condenser 
pressure of 1.1 in. of mercury. The circulating water enters the condenser at a tempera- 
ture of 60 deg. fahr. and leaves at a temperature of 77 deg. fahr. The condensate temper^ 
ature is 80 deg. fahr. The condenser is of a modern dry-tube type. The velocity of the 
cooling water is assumed to be about 5 ft. per sec. What is the required area of tube 
gurface? 



DIVISION 10 
METHODS OF RECOOLING CONDENSING WATER 

385. Condensing-Water For Steam And Ammonia Con- 
densers May Be Used Over And Over Again if some means for 
re-cooling it economically is available. Re-cooling of con- 
densing water may be necessary when, due to material limita- 
tions, or for economic reasons, an ample supply of the water 
is unavailable. Re-cooling of the water conserves the water. 

386. The Cooling Effect Of Cooling Ponds, Sprays and 
Cooling Towers, on condensing water is due to three causes. 
(1) Evaporation of the water. (2) Direct heat transfer by con- 
duction and convection, which is of minor consequence as 
compared to that of the evaporative effect. (3) Direct heat 
transfer by radiation which also is of minor consequence. The 
cooling effect of evaporation is due (See the author's Practical 
Heat) to the fact that whenever a liquid evaporates — when 
it is transformed into a vapor — an amount of heat equivalent 
to its latent heat of vaporization must be absorbed by it to 
effect the vaporization. In the atmospheric cooling of 
condensing water, practically all of this heat which is required 
to effect the evaporation of the condensing water is abstracted 
from the unvaporized portion of the condensing water itself. 
Thereby the remaining portion of the water is cooled. A 
minor portion of this heat which is required to effect the 
evaporation is abstracted from adjacent air and objects. 

Note. — The cooling effect of the direct heat-transfer (See the author's 
Practical Heat) is usually of minor consequence; see Note under 
Sec. 399. This direct-heat-transfer cooling effect is caused by the heat 
in the condensing water being conducted and radiated into the surround- 
ing air and objects. 

387. Atmospheric Recooling Of Condensing-Water, after 
the water has been discharged from the condensers, may be 
promoted by bringing the water into intimate contact with 
the air of the atmosphere and by the evaporation of a part 

329 



330 



STEAM POWER PLANT AUXILIARIES [Div. 10 



of the water. Intimacy of contact with the air and ample 
surface to promote effective evaporation is secured by break- 
ing up the mass of water into a fine spray or into a multitude 
of tiny streams or rivulets or by spreading it out over an 
extensive area in a shallow pond. The recooling effect 
depends upon: (1) The temperature-difference between the 
water and the air. (2) The relative humidity of the air. High 
humidity and high air-temperatures are drawbacks to satis- 
factory re-cooling. (3) The degree of contact-intimacy. 



Distributer-*^ ^—Lotddzv 

^-—*\-Cool!ng 
"Tower 




•■Distributer 



Motor-Driven Centrifugal! 
Circulating Pump •' 

Fig. 312. — "Burhorn" Metallic Tower 
For Cooling The Water For Double- 
Pipe Ammonia Condenser. 



Cooling- 
Tower 



<-Laololer 




Fig. 313. — "Burhorn" Metallic Tower 
For Cooling The Water For An Atmos- 
pheric Ammonia Condenser. 



Note. — Profitable operation of an atmospheric recooling system is, 
in general, mainly dependent upon the degree of effectiveness with which 
all parts of the water are brought into contact with the air. 

Notes. — Modern-Type Steam-Condensers Ordinarily Operate 
(Sec. 328) With Vacua Ranging From About 26 In. To 28 In. of 
mercury column. Assuming the temperature of the water entering 
the condenser to be about 85 deg. fahr., the discharge temperature, 
corresponding to the vacua above noted, would range from about 90 to 
110 deg. fahr. 



Sec. 388] METHODS OF RECOOLING CONDENSING WATER 331 



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332 



STEAM POWER PLANT AUXILIARIES [Div. 10 



The Temperature Of The Water Leaving An Ammonia Condenser 
(Figs. 312 and 313) of the submerged type may be from about 75 to 
80 deg. fahr., while the water from an atmospheric ammonia-condenser 
may during the summer months have a temperature of from about 75 to 
85 deg. fahr. This subject is discussed in the Author's Mechanical 
Refrigeration. 

The Temperature and Relative Humidity Of The Atmospheric 
Air (Table 388) are dependent upon the locality and the season of the 
year. For practical purposes, the local weather bureau reports may be 
referred to for this information. But where these are not obtainable, 
the relative humidity (see the author's Practical Heat) must be deter- 
mined (Sec. 389) by the use of instruments made for the purpose. 

389. To Determine The Relative Humidity Of The Air, a 

sling psychrometer (Fig. 314) may be used. This instrument is 
formed with two ordinary thermometers. The dry bulb of one 
Ti, is dry and bare, so as to be exposed 
directly to the temperature of the air. 
The wet bulb of the other, T 2 , is covered 
with cotton gauze or cloth which is satu- 
rated with water. 



Explanation. — If air is blown over the two 
thermometers, or if they are swung by rotating 
the handle, H, rapidly through the air, the one 
having the "wet bulb" will generally show a 
lower reading than the one with the "dry bulb." 
This is due to the fact that, in general, atmos- 
pheric air is not fully saturated (see the author's 
Practical Heat). It will still have some 
capacity for absorption of moisture. Therefore 
it will absorb moisture from the wet gauze which 
envelopes the "wet bulb." A cooling effect, due 
to evaporation of the moisture in the gauze, is 
thereby produced. 



Supporting Eye 




'•-Moistened- 
Cheese Cloth 



Fig. 314. — Sling Psy 
chrometer For Determin 
ing Relative Humidity. 



390. The Relative Humidity Of The 
Air Is A Function Of The Temperature - 
Difference Indicated By The Wet- And Dry-Bulb Ther- 
mometers Of A Sling Psychrometer (Fig. 314). Hence, when 
the temperature-difference shown thereby is known, the 
corresponding relative humidity may be computed therefrom, 
or it may be obtained directly from the results of such com- 
putations, which are given in Table 393. How these relative- 



Sec. 391] METHODS OF RECOOLING CONDENSING WATER 333 

humidity values are utilized in practical computations will be 
hereinafter explained. 

Note. — When there is no difference (Table 393), between the readings 
of the wet- and dry-bulb thermometers (Fig. 314), then the air is fully 
saturated with moisture. That is (Sec. 387), the air has absorbed as 
much water as it can possibly retain, at the given temperature, in a 
vaporous condition. Hence, no cooling effect, due to evaporation from 
the wet bulb, can result. The relative humidity is then 100 per cent, 
(see the author's Practical Heat). 

Example. — When the dry-bulb thermometer (Fig. 314) reads 70 deg. 
fahr., and the wet-bulb thermometer reads 60 deg. fahr., the temperature- 
difference = 70 — 60 = 10 deg. fahr. The corresponding relative humid- 
ity, from Table 393, is 55 per cent. This value is found in the same 
horizontal column with the given value, 70, of the air-temperature and 
in the same vertical column with the computed value, 10, of the tempera- 
ture-difference. 

391. The Limit Of Atmospheric Cooling Is The Wet-Bulb 
Thermometer Temperature. — Careful investigation proves 
that this is the lowest temperature attainable by cooling in 
free contact with the atmosphere (Cooling Tower Company). 
This temperature is, then, a measure of the efficiency of any 
atmospheric-cooling device. Perfect apparatus, that having 
an efficiency of 100 per cent, would reduce the temperature 
of the cooled water to that of the wet bulb. The number of 
degrees temperature decrease thus effected, would be the 
ideal range. The number of degrees temperature decrease 
attained in practice is the actual range. Hence: Actual range 
-5- Ideal range = Efficiency of the apparatus, or the percentage 
of the ideal which is actually realized. See Sec. 392 for the 
formula which expresses this relation. 

Note. — The wet-bulb temperature, therefore, bears the same relation 
to atmospheric cooling that the barometic height does to condenser 
vacua. It is the ideal minimum temperature which can be approached 
infinitely close but which can never be passed. How nearly this ideal 
minimum temperature may be attained is determined by: (1) Water dis- 
tribution. (2) Cooling surface. (3) Air supply. Increasing the effec- 
tiveness of any or all of these elements decreases the: first cost, operating 
expense, and maintenance expense. There is then, a certain degree of 
attainment toward the ideal past which it does not pay — in dollars and 



334 STEAM POWER PLANT AUXILIARIES [Div. 10 

cents — to proceed. The determination of this "point of maximum econ- 
omic effectiveness" is a problem for specialists. 

392. The Efficiency Of Any Atmospheric Cooling Device, 

cooling pond, spray nozzle installation or cooling tower, may 
be computed by the following formula : 

(92) E = 100 % n ~ ^ /2 (per cent.) 

J- fl ~ J- fw 

Wherein E = the efficiency, in per cent. T n = the tempera- 
ture, in degrees Fahrenheit, of the water coming to the cooling 
device. 7 n = the temperature, in degrees Fahrenheit, of 
the cooled water leaving the device. T/ w = the wet-bulb 
temperature of the surrounding atmosphere in degrees Fahren- 
heit, corresponding to the given relative humidity, as com- 
puted from Table 393. 

Example. — The temperature of the water entering a cooling-tower is 
108 deg. fahr. The temperature of the water leaving the tower is 88 deg. 
fahr. The temperature and relative humidity of the outside air are, 
respectively, 70 deg. fahr. and 50 per cent. What is the efficiency of 
the tower? 

Solution. — By Table 393, the difference between a dry-bulb tempera- 
ture of 70 deg. fahr. and the corresponding wet-bulb temperature, for 
51 per cent, relative humidity, is 11 deg. fahr., while the difference for 
48 per cent, relative humidity is 12 deg. fahr. Therefore, the wet-bulb 
temperature corresponding to 50 per cent, relative humidity = 70 — {ll + 
[(12 - 11) -T- (51 - 48)]} = 68.7 deg. fahr. Then, by For. (92), the 
efficiency of the tower = E = 100[(7Vi - T f2 )/(T f i - T fw )] = 100 X 
[(108 - 88) + (108 - 68.7)] = 51 percent. 



Sec. 393] METHODS OF RECOOLING CONDENSING WATER 335 



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336 



STEAM POWER: PLANT AUXILIARIES [Div. 10 



394. The Weight Of Water Vapor Which Is Contained In 
A Cubic Foot Of Atmospheric Air Is Determined By The Tem- 
perature And The Relative Humidity Of The Air. — The graph 
Fig,. 315 indicates the relation between temperature and weight 

for air of 100 per cent, relative hu- 
midity. To obtain the weight at 
any relative humidity other than 
100 per cent., multiply the value 
taken from the graph by the 
known relative humidity expressed 
decimally. 




20 40 60 80 100 120 
Air Temperature in Degrees Fahrenheit 



Fig. 315. — Graph Showing Re- 
lation Between Temperature And 
Weight Of Water Vapor In 1 Cu. 
Ft. Of Air Of 100 Per Cent. Rela- 
tive Humidity. 



Example. — The temperature of a cer- 
tain volume of air is 100 deg. fahr. Its 
relative humidity is 55 per cent. What 
is the weight, per cubic foot, of its 
moisture content. Solution. — From 
the graph of Fig. 315, the weight of the water-vapor content in 1 cu. 
ft. of air, at 100 deg. fahr. and at 100 per cent, relative humidity, is 0.003 
lb. Hence, for 55 per cent, relative humidity, the moisture content of 
the air = 0.003 X 0.55 = 0.00165 lb. per cu. ft. 

Note. — When air is "saturated," its relative humidity is then 100 
per cent, and the weight of water-vapor content in it is a maximum. 
Hence, for saturated air, the weight of its water-vapor content is deter- 
mined solely by the temperature. Like- 
wise, assuming any constant relative ~ ]2 
humidity, the weight of the water vapor g 10 
content will be determined solely by the I 8 

temperature. *£ 6 

£4 



fe 





nTnJrjfJTif ^ 




















































\w 














20 
Temp. 


40 
of 


60 60 
Air in D« 


100 120 140 
gfrees Far 


1 
re 


1 
nhe 


30 
t 



395. The Water Vapor Pressure 
Exerted By Water Vapor In Air Is 
Determined By Its Temperature 
And By The Relative Humidity — 

Water-vapor-pressure values are 

used in computing the effectiveness 

of cooling ponds and towers and 

similar condensing -water -cooling 

arrangements. The graph of Fig. 316 shows the relation 

between temperature and vapor pressure for saturated-air 

water vapor, that is, for the water vapor in air which is of 100 

per cent, humidity. 



Fig. 316. — Graph Showing Re- 
lation Between The Temperature 
And Vapor Pressure Of Saturated 
Water Vapor (Or Of Water Vapor 
In Air Of 100 Per Cent. Hu- 
midity). These Are Merely Val- 
ues Plotted From A Steam Table. 



Sec. 396] METHODS OF RE COOLING CONDENSING WA TER 337 

Note. — To Obtain The Water-Vapor Pressure Exerted By 
Vapor In Un-Saturated Air, multiply the pressure value (taken from 
Fig. 316) which corresponds to the known temperature, by the relative 
humidity, expressed decimally. 

396. Three Principal Devices For Bringing The Water And 
Air Into Intimate Contact In A Recooling System are com- 
monly available. These are: (1) The simple cooling pond or 
tank (Fig. 317). (2) The spray-fountain (Fig. 318). (3) 
The cooling-tower (Fig. 319). Each of these devices has its 
particular field of application, as will be shown in following 
Sees. 




Fig. 317. — Diagrammatic View Of A Typical Cooling Pond. (A ditch may be sub- 
stituted for the trough T). 



397. Cooling Ponds May Satisfy The Requirements Of A 
Recooling System Where Ample Ground Space Is Available. 

The operation-expense of a simple cooling pond is very low. 
The power-cost may be, and often is, practically zero. Gen- 
erally, however, for plants exceeding about 1,000-h.p. capacity, 
the area and investment necessary for an adequate cooling 
pond would be so extensive that the annual cost of the pond 
would be prohibition. Hence, for the larger plants, more 
compact devices, as spray fountains and cooling-towers, may 
be more economical and satisfactory. 

398. The Rate Of Evaporation From A Simple Cooling 
Pond, When The Air Is Perfectly Calm, may be computed 
by the following formula : 



(93) W = (243 + 3.7T/)(P„ 

22 



P V M) ( grains per sq. ft. perhr.) 



338 



STEAM POWER PLANT AUXILIARIES [Div. 10 



Wherein W = the weight of water evaporated, under calm 
air, in grains per square foot per hour; it may be increased 
materially by the effect of wind; see following note. Tf = 
the temperature of the water, in degrees Fahrenheit. P v = 




w^m 



; I-E leva t- i o n -v.." »•".„* •*/•.'■ £/*. '» '.".' 



Fig. 318. — Diagram Showing Schutte And Koerting Double Spraying System. In 
Winter One Side Is Shut Down. (Spray nozzles are set at an angle of 45 degrees to 
the horizontal.) 



the vapor pressure, in inches of mercury-column, as taken 
from the graph (Fig. 316), for saturated air at the given 
temperature. M = the per cent., expressed decimally, of 
the relative humidity of the air, as found in Table 393. 



Sec. 398] METHODS OF RECOOLING CONDENSING WATER 339 

Note. — In the expression " (P v — P V M)," in the above equation, il P v " 
is the vapor pressure which would be exerted by a saturated air vapor 
at the given temperature and "P V M" is the vapor pressure actually- 
exerted by the non-saturated air vapor under consideration. Their 
difference is a measure of the tendency to promote evaporation. See 
Sec. 395. 



Tile 
{Filling 




•' ' SUcflon p 'Pe >'■■■ Lower Overflow Pipe Closed -"" - v •*. : "■"*.*• • [ m<ir overflow Pipe Open 
I-Co)d-Well Level Raised To Permit I-Cofol-WeN Water Level Lowered For 

Forced Draft Operation Natural Draft Operation 

Fig. 319. — Worthington Cooling-Tower Using Either Forced Or Natural Draft. 



Note. — In Practice, The Weight Of The Evaporation may, due 
to normal wind velocities, be from 2 to 12 times greater than the value 
obtained by the preceding formula. A fair average is from 6 to 8 times 
the computed value. 

Note. — For. (93) may be rewritten as follows: 

(243 + S.7Tf)(P v - PvM) 



(94) 



W, 



"000 



(lb. per sq. ft. perhr.) 



Wherein W w = weight of water evaporated, under calm air, in pounds 
per square foot per hour. T/, P v and M are as given in For. (93). 



340 STEAM POWER PLANT AUXILIARIES [Div. 10 

Example. — The temperature of the water in a cooling pond is 80 deg. 
fahr. The air-temperature is also 80 deg. fahr. The relative humidity- 
is 75 per cent. Assuming that the prevailing wind-velocity multiplies, 
8 times, the rate of evaporation under calm air conditions, what is the 
approximate rate of evaporation, in pounds per square foot per hour? 
How many square feet of the pond surface are required to give off each 
pound of the evaporation? Solution. — The graph of Fig. 316 shows 
the water vapor presswe corresponding to a temperature of 80 deg. fahr. = 
1.0 in. of mercury-column. By For. (93), the rate of evaporation in calm 
air = W = (243 + S.7T f )(P v - P V M) = (243 + 3.7 X 80) X [1 - (1 X 
0.75)] = 134.75 grains per sq. ft. per hr. The avoirdupois pound con- 
tains 7,000 grains. For the prevailing wind-velocity, the approximate 
rate of evaporation = (134.75 X 8) -5- 7,000 = 0.154 lb. per sq. ft, per hr. 
Therefore, 1 -r- 0.154 = 6.49 sq. ft. per lb. per hr. = number of square 
feet of pond surface necessary to evaporate 1 lb. of water per hour. 

399. The Evaporation Of One Pound (1 lb.) Of Condensing 
Water Is Equivalent To The Abstraction Of About 1,000 B.t.u. 
From The Water. — This is true because, as shown in any- 
steam table, the latent heat of vaporization (or evaporation) 
of water vapor is, for the vapor pressures encountered in 
cooling-pond, spray-nozzle and cooling-tower practice, about 
1,000 B.t.u. The approximate vapor pressure in any instance 
is that, as taken from the graph of Fig. 316 corresponding to 
the existing air temperature. The exact vapor pressure is 
that taken from Fig. 316 for the given air temperature, multi- 
plied by the relative humidity. The relative humidity may 
be obtained as explained in Sec. 390. 

Example. — If 8 lb. of water evaporates from the water in a cooling 
pond, a spray pond or a cooling tower, there will thereby be abstracted 
from the water in the pond approximately: 8 X 1,000 = 8,000 B.t.u. 

Note. — With low air-temperatures, radiation of heat from the pond, 
and transfer of heat to the air by conduction and convection, assist, to 
some extent, in cooling the water. The loss of heat by evaporation may 
under this condition, be somewhat reduced. With high air-temperatures, 
the reverse is true. It is generally observed that, in moderately -warm 
weather and under ordinary conditions, approximately 90 per cent, of 
the cooling effect is due to evaporation and 10 per cent, to other causes. 

400. In Estimating The Requisite Surface Area For A 
Simple Cooling Pond, for cooling the circulating water of a 
steam condenser, (Fig. 320) it may, safely, be assumed that: 
(1) The total heat given up or lost by the cooling-pond water is 



Sec. 400] METHODS OF RECOOLING CONDENSING WATER 341 



solely that abstracted from the water by evaporation. (2) The 
total heat imparted to the water is the heat given thereto in the 
condenser by the steam during its condensation therein. Hence, 
if the temperature of the condensing water in the pond is to 
be maintained constant, the pond area must be sufficiently 
great that it will, by its evaporative effect, release the same 
amount of heat per hour as is imparted to it per hour by the 
condensing steam. Now, it may also be assumed that the 




I-With Jet Condenser [Circulating Water Suitable for Boiler Feed)': 

Cooling Pond* 




'Meter 

HE- With Jet Condenser {Circulating Water Unsuitable for Boiler Feed) 
Fig. 320. — Three Possible Arrangements Of Condensing Equipment. 

heat, in B.t.u., which is given up to the condensing water by 
1 lb. of steam when the steam is condensed, is equal to the 
heat, in B.t.u., which is abstracted from the cooling-pond 
water by 1 lb. of the water when it evaporates. In both cases, 
the amount of heat is about 1,000 B.t.u.: 

Therefore: The approximate requisite total pond-area ivill 
result if the area (Sec. 398) which is required to give off lib. of 
evaporation per hour, is multiplied by the number of pounds 
of steam which is condensed per hour. 

Note. — The Above Assumptions Are Not Strictly Accurate. 
But inasmuch as the resulting values which are obtained by using For. 



342 STEAM POWER PLANT AUXILIARIES [Drv. 10 

(93) and (94) must be increased (Sec. 398) by from 2 to 12 times to cor- 
rect for wind effect, the above-indicated method is sufficiently accurate 
for estimating. 

Note. — The Cooling Effect Of The "Make-Up" Water May Be 
Disregarded because the make-up water — that which must be replen- 
ished because of evaporization, windage and other losses — is less than 
2 to 3 per cent, of the total amount of water which is circulated. The 
variation in evaporative effect due to wind (Sec. 398) will more than 
offset the cooling effect of the make-up water. 

Example. — A steam condenser (any type) is required to condense 
3,600 lb. of exhaust steam per hour. The water will be discharged to a 
cooling pond for recooling. The temperature of the discharge water 
will be 80 deg. fahr. The air temperature is also 80 deg. fahr. The 
relative humidity is 75 per cent. What should be the area, in square 
feet, of the cooling pond? 

Solution. — It is computed in the Example under Sec. 398 that, under 
the conditions just specified, each 6.5 sq. ft. of pond area will evaporate 
1 lb. of water per hour. Hence (Sec. 400: Total pond area = Area re- 
quired to give of 1 lb. evaporation per hour X Number of pounds steam 
condensed per hour), the total pond area should be: 6.5 X 3,600 = 23,400 
sq. ft. 

401. The Requisite Area For A Simple Cooling Pond 
Cannot In Any Case Be Computed Precisely, due to the 

numerous variables which are involved. The most important 
of these are (1) the temperature and humidity of the air (Sec. 
387) (2) the solar reheating effect and, particularly (3) the wind- 
velocity (Sec. 398). The type of condenser and the kind of 
condenser-service, whether steam or ammonia, may also 
affect the problem. 

Note. — When A Cooling Pond Is Located On The Roof Of A 
Building, There Is More Solar Reheating. Hence, in such loca- 
tions, the pond area must, usually, be greater for an equivalent effect. 
If spray nozzles are used over a roof pond, the same number spaced in 
the same way will not, ordinarily, give as good results as over a surface 
pond. 

402. Some Cooling -Pond -Area Data Are: It has been 
determined (Proceedings A. S. M. E., Apr., 1912, page 607) 
that, in the northern part of the United States, 120 sq. ft. of 
cooling-pond surface will suffice for 1 h.p. of steam-condenser 
service. This value is based upon a 26-in. vacuum and a 
steam consumption of 15 lb. per h.p. per hr. Cooling-pond 
area is sometimes determined upon the assumption that 8 sq. 



Sec. 403] METHODS OF RECOOLING CONDENSING WATER 343 

ft. will suffice for each pound of steam condensed. Also, 
that 1 sq. ft. of pond area will give off 4 B.t.u. per hour per 
deg. fahr. difference between the water-and air- temperature 
in summer, and 2 B.t.u. in winter. 

403. The Depth Of Simple Cooling-Ponds is usually from 
3 to 4 ft. The depth has little influence on cooling effective- 
ness, provided the surface-area is ample, since the cooling is 
determined almost wholly by the surface area which is exposed 
to the air and from which evaporation can take place. 

404. Spray Fountains Are Often Used In Connection 
With Cooling-Ponds.— This arrangement (Fig. 318 and 321) 




Sprays 
Concrete Pier. 



-•'•:" '■ ) : [ ' [ ':i'-A -.Uii ii'iH I'jh tfM ili.l 



Discharge 

Noizles P i peFrom 

_J^ Condenser 



Fig. 321. — -Spray Pond With Cooling-Tower Company's "Impact" Nozzles. (Space 
between sprays permits effective air circulation.) 

permits of a considerable reduction in the area of the pond. 
The fine division of the water particles by the sprays (Fig. 
318) insures a maximum of water surface in small space and 
thereby facititates the cooling effect due to evaporation and 
air contact. 



Note. — The passage of the water through the cores of the nozzles 
(Figs. 322 and 323) on a spray fountain, breaks it (Figs. 324, 325 and 326) 
into a fine mist. These nozzles are, generally, set either vertically 
(Fig. 324), at an angle of 45 deg. with the horizontal (Fig. 318), or at an 



344 STEAM POWER PLANT AUXILIARIES ' [Div. 10 

angle of 60 deg. with the horizontal (J, Fig. 327). These arrangements 
secure a wide distribution of the spray. They also tend to induce air- 
currents, even on calm days, which greatly augment the cooling effect. 



■Outlet 




Sto/no/oircl Pipe Thread, UsuallylJTo J //?.-"-> 



Fig. 3 2 2.— Badger Spray Nozzle. 
(Badger & Sons Co., Boston.) 



Outer 
Nozzle 



< -Inner 

Nozzle 




^-Stornolotro! Pipe Thread, Usually If To 3 In. 
Fig. 323— Koerting Multi-Spray Nozzle. 




Fig. 324.— Form Of Spray From Single Spray-Nozzle. 

405. The Conditions Which Mainly Control The Amount 
Of Recooling Produced By Spray Fountains have been deter- 
mined by tests. It has been demonstrated (Fig. 328): (1) 



Sec. 405] METHODS OF RECOOLING CONDENSING WATER 345 

That recooling is more affected by the air-temperature and 
humidity than by the temperature of the water coming from the 






^■■Spmy..^,^: 



m?m* 









Pi'pe- 



Fig. 325. — Form Of Intermingled Spray From Three Nozzles. 

condensers. (2) That with 80 to 90 per cent, relative humidity, 
the water-temperature can be lowered to within 12 or 13 deg. 
fahr. of the dry-bulb air-temperature. (3) That with 20 to 30 



^:X\\\\\\! ']!///, 

'■-. n\F.6i n-5 hoi peoT- -. 

Cone- 
Shapeol , T VX ^</ -Cas/- 
Cup-'' J?/ X V:' Bronze 
£l-~ 4$^- - ; ^\ Nozzle 

Two---, 
Converyn 

and Partly-\^p^rpy[ Pipe 
!ntersecting\-^--— --U Connection 

Jets ' ' 

Fig. 326. — Cooling Tower Company's Impact Spray Nozzle. (Designed to minimize 
possibility of clogging and to promote air circulation.) 

per cent, relative humidity, the water-temperature can be lowered 
about 8 deg. fahr. below the dry-bulb air-temperature. (4) That 
the loss of water is usually about 6 per cent. 




346 



STEAM POWER PLANT AUXILIARIES [Div. 10 



Atmospheric 
Exhaust Head-.. 







Fig. 327. — Utilizing Roof-Space For Spray-Cooling. (Koerting.) 



2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 52 34 36 je 40 42 44 46 TEST NO. 

233444445 5 5 6 6 8 8 8 8 9 9 9 10 U U PAY 0T MONTH 
430 1100 7.30 10.00 12.00 300 5.00 6.00 10.0012.00 3.00 230 1030 10.30 2.30 4.30 7.30 1000 12.00 230 9.00 2.00 4.00 HOUR 0T PAY 
T.M AH ?rt. A.M. NOON ?.M. A.M. BR AH NOON PH. AH A.H AM RH, PH PM. AH NOON EH AH. P.M. BH 



Water On 




N NW NW NW NW NW NW NW W NW NW 5W W W NW NW W W W NW SONW .... 

Y — Cloudy "'■■Cloudy Cloudy— ■■ -Ram 

Clear Except As Noted On The PfiWAnd //•-*■ 



Fig. 328. — Graph Of Spray-Nozzle Tests On A " Cooling-Tower-Company " Installa- 
tion At Silver Springs, New York, September, 1919. (Flat spray nozzles were used. 
Circulation = 5,000 gal. per min. on steam-condenser duty. The capacity (size) of 
these nozzles was 60 gal. per min. at 6>£ lb. per sq. in. pressure. The value of " K n " 
used in the guarantee equation, For. (95), was "5.7.") 



Sec. 406] METHODS OF RECOOLING CONDENSING WATER 347 

406. To Compute The Temperature Reduction Which Can 
Be Effected By A Spray-Nozzle Installation the following 
formula can be used. It is quoted from The Cooling Tower 
Company's Catalogue and is the basis of its guarantees. 



(95) T /2 = 



{ (7V +460) + (7V + 460) J . _ ^ + m) - 
K n X 100,000,000 



Wherein, all temperatures are in degrees Fahrenheit and; 
T/2 = temperature of cooled water after spraying. T/i = 
temperature of water before spraying. T/ x = (4:Tf w + T fd ) 
-f- 5. T/d = dry-bulb-thermometer or air temperature. T/ w 
= wet-bulb-thermometer temperature. K n = a constant = 
5.1 for average installations operating at 6J^ lb water pres- 
sure but it may vary from 4.0 to 5.7. These values for K n 
were determined from tests made by the Cooling-Tower 
Company using the impact nozzle of Fig. 326. K n varies 
with the type, size and spacing of the nozzles and with the 
water pressure and wind velocity. For equal operating pres- 
sures and atmospheric conditions, the value of K n depends 
mainly on the pond exposure, the size of the nozzles and the 
ratio of pond area to water sprayed. 

Note. — The Predetermination Of The Proper Value For K n , 
for any given installation, requires extensive experience in this branch 
of engineering. Consequently, to design a cooling system which will 
develop a given value of K n , a thorough knowledge of the local conditions 
is necessary as well as a practical understanding as to the effects of such 
conditions. It is feasible, should the service conditions justify the expen- 
diture, to so design the system that the value of K n will be as low as 4.0 
or even less. 

Example. — See Fig. 328 which indicates the approximate agreement of 
of actual observed values with values obtained by computation with 
For. (95) using a value of 5.7 for K n . 

Note. — By using values from Table 388, the probable temperature 
reduction which may be expected in any locality can be computed. 

Note. — Performance Guarantees On Combined Condenser-And- 
Spray-Cooling Outfits can be obtained from certain manufacturers — 
Schutte & Koerting Co. for example. In such guarantees, the vacuum 
performance of the condenser is based on the outside-air temperature — 



348 



STEAM POWER PLANT AUXILIARIES [Div. 10 



not on the temperature of the injection water; a standard relative humid- 
ity as is assumed. 

407. The Size And Number Of Nozzles To Be Used In A 
Spray-Fountain (Table 408) depends upon the quantity of 
water to be handled. It is commonly assumed that a single 
spraying system will, under normal conditions, cool the water 
about 20 to 30 deg. Fahr. This is considered sufficient 
(Table 410) for ordinary steam-condenser service. However, 
it is often considered desirable to spray from 25 to 55 per cent, 
of the condensing water a second time before sending it 
through the condenser. 

408. Table Showing Spray-Nozzle Capacities In Gallons 
Per Minute. (Schtjtte & Koekting Company). 

Note. — Nozzles of 2-in. pipe-size are most frequently used. These are 
commonly regarded as the most economical. The outlet orifice in the 
tip of a 2-in. nozzle has a diameter of about 0.8 in. The hydraulic 
pressure required to force the water through the nozzles should never 
exceed about 14 lb. per sq. in., gage. 



Pipe-size of 
nozzle, in inches 


Pressures on nozzles, in pounds per square inch 


5 


6 


7 


8 


9 


10 


2 


54 


60 


65.5 


70.5 


75 


78 


2M 


77 


85 


92 


98 


103 


106 


3 


115 


125 


133 


140 


146 


151 



409. The Spacing Of The Nozzles In A Spray-Fountain 

depends mainly upon the design and size of the nozzles. 
Centrifugal nozzles of 2-in. size are usually spaced about 8 to 
10 ft. from center to center. Nozzles of larger size may be 
set proportionately further apart. 

Note. — A typical installation, spraying 4,800 gal. per min., consists of 
9 rows of nozzles, with 8 nozzles in each row. Thus, each nozzle sprays 
4,800 t (9 X 8) = 66% gal. per min. The rows are 20 ft. apart, and the 
nozzles are spaced 13 ft. between centers. A 2-in. nozzle (Table 408) at 
a little over 7 lb. per sq. in. water pressure would meet these requirements. 



Sec. 410] METHODS OF RECOOLING CONDENSING WATER 349 



< 

S 
o 
O 

o 

i— i 

S3 

H 
O 

W 

H 
o 

CO 



o 

g 

CO 



n 

I 

J 

CO 

.2 

S 



Excess of. 
water temp. 

above air 
temp, of 70 

deg. fahr. 


<N 


CO 


- 


Temp, of 
water after 
spraying in 

deg. fahr. 






oc 


Cooling 

obtained 

in deg. 

fahr. 


o 


OS 


OS 
CO 


Cooling 

required 

in deg. 

fahr. 


O 


as 


CO 


Temperature 
of water leav- 
ing condenser 
in deg. 
fahr. 


OS 


o 


c 

1—1 


Rise in con- 
denser of con- 
densing- 
water tem- 
perature in 
deg. fahr. 


O 


OS 


CO 


Gallons of 

condensing 

water per 

pound of 

steam 


O 




co 

CO 


Pounds of 

condensing 

water per 

pound of 

steam 


O 


CO 


o 

CO 


Condenser- 
vacuum 
in inches 


00 




<o 



350 STEAM POWER PLANT AUXILIARIES [Div. 10 

411. The Ground -Area Required For Spray -Fountain Ponds 

is much less than that required for simple cooling-ponds. It is 
commonly assumed that for small plants, under 500 h. p., 1 sq. 
ft. of surface will suffice for the cooling of 150 lb. of water per 
hr. For plants of about 5,000 h.p., 1 sq. ft. of surface will 
usually suffice for the cooling of 250 lb. of water per hr. For 
plants of about 1,000 h. p., 1 sq. ft. of surface may be assumed 
as sufficient for the cooling of 200 lb. of water per hr. 

Note. — Spray-Fountains Should Be Surrounded By Wind- 
Breaks, in order to avoid excessive water-loss, due to heavy winds. 

412. Spray-Fountains Are Sometimes Located On Power- 
House Roofs (Fig. 327). This is usually done where ground- 
area is unavailable. The extra power required for elevating 
the condensing water may, with some types of condenser 
installations (Fig. 327), be offset by utilizing the hydrostatic 
head, thus obtained, for sending the recooled water through 
the condenser. 

Explanation. — The water from the hot-well (Fig. 327) is pumped to 
the spray-nozzles J, by the centrifugal pump, G, through the discharge- 
pipe, H. The recooled water in the spray, V, runs into the trough K. 
It then flows through the low-level-jet eductor condenser C, under the 
head due to its elevation above the hot-well. 

413. The Power Required To Operate A Spray-Fountain 

in connection with a steam-condenser equipment is not great. 
The pressures generally used seldom require more than 1.5 to 
2 per cent, of the main-engine power. This is equivalent to 
about 10 per cent, of the power saved by condensing over non- 
condensing operation. Very often, the circulating pump (Sec. 
353) may be used to deliver the water directly to the nozzles. 
Installation of additional pumping equipment is thereby 
avoided. This can generally be done with surface-condensers 
(Sec. 335) and low-level jet- or eductor-condensers (Sec. 336) 
but not (Sec. 339) with barometric condensers. 

414. A Cooling -Tower (Fig. 319) consists, essentially, of a 
tall, narrow, wooden or sheet-iron structure, so arranged 
internally that after the warm condensing-water has been 
elevated to the top under pump-pressure, it will fall, by gravity, 
in a multitude of thin sheets or trickling streams, into a reser- 



Sec. 414] METHODS OF RECOOLING CONDENSING WATER 351 
voir or sump, S, which is located beneath the tower. In falling 



Notches-,-—^ 





Cypress ; , 
Planks-'"' 



Water Supplied H< 



Cypress Trouaf?-. 



Secondary Breaking/ Up , Cypress Triangular 
•■ of Water 'Particles- \ Borfflts-., 

* ^ ) ! .' r xK ' | / (^w\ y ffi] ) 

, : ' r#-": i : f#*. '£ :^Tl i , ;. 



Fig. 329.— Cypress Board Checker Fig. 330.— "Wheeler" Cooling-Tower Splash 
Work For Cooling Towers. Counter-Flow System. 



k ^5wamp Cedar Lumber 



Separator 




ZZZ 



D • Section Through Distributor kno) Decks 



Fig. 331. — Distributor And Decks Of "The Cooling-Tower Company' 1 Design. (Each 
tower contains 10 intermediate decks arranged one above the other. ) 

it is cooled by the air which surrounds it. The devices for 
dividing the water into fine sheets, droplets or sprays may 



352 



STEAM POWER PLANT AUXILIARIES 



[Div. 10 



consist of: (1) Checker work (Fig. 329). (2) Corrugated sur- 
faces. (3) Troughs or baffles (Fig. 330 and 331) . (4) Galvanized 
steel wire screens or perforated trays. (5) Masses of tile-tubing 
or galvanized iron pipes (Fig. 332) set vertically. 



■Interlocking Pipes*., 




Air And Water-:" 
Spaces, 

Fig. 332. — Interlocking Pipe Filling In Mixing Chamber Of Worthington Tower. 

Note. — In every case, the tower is open at the top, and is so arranged 
at the bottom that atmospheric-air will circulate (either by natural draft 
or by pressure of a fan-blower) through the descending water. The 
water gives up its heat to the ascending air-currents by evaporation, 
convection and radiation. 

Perforated Pipe Distributer-^ i 0UYre5 . 
Trough , 




'-Collecting Pan 
Fig. 333. — "Burhorn" Open Or Atmospheric Cooling Tower. (Louvres removed 
from one side to show construction. In some cities wooden cooling towers are pro- 
hibited because of fire risk.) 

Note. — From 75 To 85 Per Cent. Of The Recooling Effected 
In A Cooling-Tower Results From Evaporation in most power 
plant installations. The percentage of recooling effected by conduction, 



Sec. 415] METHODS OF RECOOLING CONDENSING WATER 353 

connection and radiation, both in the tower and from tlu> piping which 
conveys the water thereto, is usually, in power plants, comparatively 
insignificant probably rarely exceeding more than 2 per cent. But where 
towers are used for cooling water from high temperature stills, where the 
cooling range may be 100 deg. fahr. or more, then, the combined cooling 
effect due to radiation and conduction may be greater than that due to 
evaporation. 




recc ee^ eccec e: 



LC.6 1 bl ^ Wat er e^^ 

Fig. 334. — Section Of Typical Wheeler- 
Balcke Natural Draft Cooling-Tower. 





^ ^b ;&::p v ? -v- s 



Fig. 335. 



-Worthington Forced Draft 
Cooling-Tower. 



Wood Checker Work (Fig. 329) For Cooling Towers usually con- 
sists of 1 X 4-in. cypress or swamp-cedar boards set on edge and spaced 
about 4 in. apart. 



415. Cooling Towers May Be Divided Into Four General 
Classes : (1) Open or atmospheric-towers (Fig. 333) using 
natural draft. (2) Closed or chimney-flue towers (Fig. 334) 

23 



354 



STEAM POWER PLANT AUXILIARIES [Div. 10 



using natural draft. (3) Closed or chimney-flue towers (Figs. 
335 and 336) using forced draft. (4) Closed or flue-towers 



mm 







K^A^WA^27^^ S ^A^^^\ ; 



pj^^i^^Msc 



Fig. 336. — Forced Draft Cooling Tower With Surface Condenser. (Worthington 

Company.) 

(Fig. 319) using either forced or natural draft. The sides of 
open or atmospheric towers (Fig. 333) are usually louvred, 
(Fig. 337) to prevent the water from being blown out of the 




Bolts- 



I- End 
Section 



I- Isometric View 
Fig. 337. — "Burhorn" Sheet-Metal Cooling-Tower Louvres. 



tower. Louvres actually decrease the cooling effect but must 
be employed to minimize water waste. 



Sec. 416] METHODS OF RECOOLING CONDENSING WATER 355 

416. The Closed Or Flue-Towers Are Completely Enclosed, 
Except At Top And Bottom. — Openings are provided in the 
base for admission of the fan blast in the one case or the 
natural air-currents in the other. Natural draft in these 
towers depends entirely upon the chimney action of the tower. 

Note. — Choice Of Forced Or Natural Draft mainly depends up- 
on space considerations on the one hand and operating-cost on the other. 
A forced-draft installation occupies less space than one using natural 
draft, but the operating expense is greater. Where cooling-towers are 
designed (Fig. 319) for using either forced or natural draft, the forced 
draft may be used during the hot season and the natural draft in cool 
weather. 

The Open Tower (Fig. 333) Permits A Somewhat Greater Loss 
Of Water Than The Closed Tower (Fig. 334). This is due to winds 
blowing through the louvres of the open tower. Generally, the air does 
not mingle so effectively with the water in open towers as it does in 
closed towers, but the closed tower must be larger for the same cooling 
effect. In many fan (forced-draft) towers, the water lost is greater than 
in an atmospheric tower of about the same size. This is because a large 
amount of water, as entrained moisture, is carried away in the forced- 
draft towers due to the high air velocity. Since such water has not been 
evaporated, it represents pure waste — it has performed no useful work 
of cooling by its evaporation. With a forced-draft and an atmospheric 
tower operating side by side, the water loss from the forced-draft tower 
may be as great as 10 per cent, and that from the atmospheric tower as 
small as 2 per cent. 

417. The Principles Involved In Cooling-Tower Computa- 
tions are similar to those pertaining to cooling-ponds and 
spray-fountains. The cooling effect depends upon the water- 
and air-temperatures, the relative humidity of the air, and the 
effectiveness of air-and-water contact. Towers of different 
types vary in the effectiveness with which the air is utilized 
as a cooling medium. 

418. Computations To Determine Cooling-Tower Per- 
formance Should Be Based On The Results Of Tests And 
Practice rather than on entirely theoretical assumptions. If 
the condition of the atmosphere as to temperature and humi- 
dity, the temperature of the water coming from the condensers, 
the quantity of water each unit-volume of air will absorb, 
and the degree of efficiency under which the tower will operate, 
are known, then reasonably-close approximations may be 



356 STEAM POWER PLANT AUXILIARIES [Div. 10 

made for any specific case by applying the general methods 
of computation (Sec. 398) previously given for cooling-ponds. 
The general method is illustrated in a following example. 

Note. — In the operation of a cooling tower, the same water is used 
over and over again. Through the process of cooling there is a certain 
loss which must be made up from some outside source. The water which 
must be supplied to compensate for this loss is known as the make-up 
water. Make-up water is equal to: (water lost by evaporation) + (water 
which is splashed or blown out of the cooling tower.) Assuming that there 
is no loss except that due to evaporation, the amount of heat (in B.t.u.) 
taken away from the water in circulation, will (See Sec. 400) equal the 
number of pounds of water lost, multiplied by approximately 1,000. 
In other words, every pound of water evaporated will carry away 1,000 
B.t.u., and cool 1,000 lb. of water 1 deg. fahr., or 100 lb. of water 10 deg. 
fahr., etc. Therefore, to cool 100 lb. of water 10 deg fahr., requires the 
evaporation of 1 lb. of water, or 1 per cent, of the amount cooled. Thus, 
theoretically, the make-up water will be 1 per cent, of the water circu- 
lated, to cool the water 10 degrees. Actual tests on several Burhorn 
towers under different conditions, have shown the actual loss to be 
less than 1% per cent, of the total amount circulated, or practically 
that due to evaporation. 

Under usual ammonia-condenser conditions, a cooling tower may be 
expected to cool the water by from 5 to 14 deg. fahr. ; about 10 deg. fahr. 
is a reasonable expectancy. For steam condensers, a tower may be 
expected to decrease the temperature of the water by from 20 to 50 deg. 
fahr. 

419. To Compute The Average Temperature Reduction 
Effected In Summer Weather By Atmospheric Cooling 
Towers now in operation in this country and abroad, use the 
following empirical formula which is derived from the results 
of a large number of tests. It is quoted from The Cooling 
Tower Company's Catalogue. 

(96) T fa = T >* + 2 l f ™ + T » (deg. fahr.) 

Wherein, all temperatures are in degrees Fahrenheit and: — 
T fa = average temperature, of the cooled water which leaves 
cooling towers. T fd = dry-bulb-thermometer or air temper- 
ature. T/ w = wet-bulb-thermometer temperature. T s\ — 
temperature of water entering the cooling tower. 



Sec. 420] METHODS OF RECOOLING CONDENSING WATER 357 

Examples. — See Tables 421 and 422 which show average values com- 
puted with the preceding formula. By using values from Table 388, 
the probable temperature reduction which may be expected in any 
locality can be computed. 

Note. — Cooling towers can be designed which will, for certain cooling 
ranges and atmospheric conditions, reduce the cooled-water temperature 
by from 10 to 50 per cent, below that given by the preceding formula. 
The possible maximum temperature reduction is determined by the 
cooling range and by atmospheric conditions. See Tables 421 and 422. 

420. Typical Data Pertaining To Cooling-Tower Perform- 
ance have been obtained from a series of tests made with 
closed cooling-towers using forced draft. They are as follows : 

Data. — Quantity of water circulated = 640 gal. per min. Tempera- 
ture of air entering the tower = 70 deg. fahr. Temperature of air 
leaving the tower = 94 deg. fahr. Relative humidity of air entering 
the tower = 50 per cent. Relative humidity of air leaving the tower = 
100 per cent. Temperature of water entering the tower = 108 deg. 
fahr. Temperature of water leaving the tower = 88 deg. fahr. quantity 
of air circulated = 50,000 cu. ft. per min. Efficiency of tower (For. 
92) =51 -per cent. These data represent about average practice for 
the given type of installation. 

Example. — Using the above data, and allowing 8.3 lb. to the gallon, 
the heat added to the water while passing through the condenser = 
640 X 8.3 X (108 - 88) = 106,240 B.t.u. per min. Assuming the 
specific heat of air to be 0.019 B.t.u. per cu. ft., the heat which the air ab- 
sorbs, by convection and radiation, from the water in the tower = 50,000 X 
0.019 X (94 - 70) = 22,800 B.t.u. per min. = (22,800 + 106,240) X 
100 = 21.46 per cent, of the heat which the water absorbed in the con- 
denser. Hence, the heat which the water gives off by evaporation — 
106,240 - 22,800 = 83,440 B.t.u. per min. = (83,440 + 106,240) X 
100 = 78.54 per cent, of the heat which the water absorbed in the con- 
denser. Assuming (Sec. 400) that each pound of the evaporation ab- 
stracts 1,000 B.t.u., the water-loss = 83,440 -^ 1,000 = 83.44 lb. per 
min. = 83.44 + (640 X 8.3) X 100 = 1.57 per cent. Wind losses might 
increase this to over 2 per cent. In practice, the usual water loss may be 
from 2 to 3 per cent. 

Note. — The Per Cent. Of Water-Loss From Cooling Towers, as 
noted above, is less than the lowest per cent, of loss that can be obtained 
with spray-fountains. This is an important item in favor of the cooling- 
tower. 



358 



STEAM POWER PLANT AUXILIARIES [Div. 10 



o w 

H £ 

o 

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O 



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CO +i 





On 


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Sec. 422] METHODS OF RECOOLING CONDENSING WATER 359 



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360 STEAM POWER PLANT AUXILIARIES [Div. 10 

423. A Method Of Computing The Proportions Of A Cool- 
ing-Tower will now be explained by the use of an illustrative 
example. Cooling-tower design is — because of the necessity 
of using results from existing installations as precedents — 
properly a function of men of considerable experience in this 
particular branch of engineering. 

Example. — A forced-draft cooling-tower is required to re-cool 1,000,000 
lb. of condensing water per hour through 25 deg. fahr. The circulating 
air is assumed to be at a temperature of 75 deg. fahr. when it enters the 
tower and at 105 deg. fahr. when it leaves. What should be: (1) The 
total horizontal cross-seciional area? (2) The total horizontal length of each 
side? (3) The total height of the checkerworkf 

Solution. — It may be assumed that the tower is to be furnished with 
a cypress-board checker-work (Fig. 329) for dividing the descending water 
into a multitude of thin sheets. Practice has shown that air velocities, 
in cooling-towers, of about 700 ft. per min. produce the best results. 
It may be assumed that the evaporating or cooling surface afforded by 
the cypress boards is about 8 sq. ft. per cubic foot of space occupied by 
the checkerwork. It may further be assumed that about 64 per cent, 
of the total horizontal cross-sectional area of the checkerwork is effective 
area, or free area. Also that 20 B.t.u. per hour, per degree of cooling, 
will be abstracted from the condensing water for each square foot of 
evaporating surface. 

The water will absorb, in the condenser approximately 1 B.t.u. per lb. 
for each deg. fahr. of temperature increase. Hence, the total quantity of 
heat to be abstracted in the cooling-tower = 25 X 1,000,000 = 25,000,000 
B.t.u. per hour. The quantity of heat abstracted per square foot of cypress- 
board evaporating surface = 20 X 25 = 500 B.t.u. per hour. Therefore, 
the requisite total area of evaporating surface = 25,000,000 -f- 500 = 
50,000 sq. ft. Hence, the total volume of space to be occupied by the checker- 
wood = 50,000 v8= 6,250 cu. ft. 

Assuming (Example subjoined to Sec. 420) that about 21.5 per cent. = 
0.215 of the heat in-the condensing water passes to the air by convection 
and radiation, the total quantity of heat so removed = 25,000,000 X 0.215 
= 5,375,000 B.t.u. per hour. Therefore, assuming the specific heat of 
air to be 0.019 B.t.u. per cu. ft., the requisite quantity of air, of the given 
entering and leaving temperature, = 5,375,000 -J- [0.019 X (105 — 75) 
= 9.429,824 cu. ft. per hr. = 9,429,824 -h 60 = 157,167 cu. ft. per min. 

For an air- velocity of 700 ft. per min., the requisite effective cross- 
sectional area of the checker work = 157,167 -f- 700 = 224.5 sq. ft. 
This being about 64 per cent. = 0.64 of the total cross-sectional area, 
the requisite total area = 224.5 ■*■ 0.64 = 351 sq. ft. Hence, the length 
of each side of the square base of the checker work = V351 = 18.7 ft., or, 
approximately, 18 ft. 8.5 in. The requisite height for the checker work, 
then, = 6,250 -*- 351 = 17.8 ft., or, approximately, 17 ft. 9.5 in. 



Sec. 424] METHODS OF RECOOLING CONDENSING WATER 361 

Note. — The Total Height Of A Cooling-Tower, of the type speci- 
fied above, would be given by the sum of the fan-height + the height of the 
checker work + 2 ft. for the height of a distributing trough (Fig. 331) + about 
4 ft. for the depth of the sump or well. If the tower is erected at the 
ground level, the sump may be sunk below the surface of the ground. 

Note. — The Height Of The Fan-Blower Required For A Cool- 
ing-Tower may be obtained from manufacturers' tables of the dimensions 
and capacities of such blowers. Typical related data pertaining to fan- 
draft towers, for use in connection with condensing-engine plants, are 
given in Table 424. 

424. Table Of Related Data Pertaining To Forced-Draft 
Cooling-Towers For Use With Condensers Of Compound 
Condensing Engines. 



Capacity 

of con- 
denser, in 
horse 
power 



Height 
of cool- 
ing tower, 
in feet 



Dimensions 
of cooling- 
tower at 
base, in feet 



Number 
and size 
in feet, 
of fans 



Speed of 
Fans in 
Revolu- 
tions per 
min. 



Power re- 
quired 
for Fan 
in horse 
power 



50 


25 


19 X 19.5 


1 - 6 


110 


1.25 


75 


25 


19.8 X 20.0 


1 - 6 


160 


1.75 


100 


25 


20.0 X 20.8 


1 - 7 


145 


2.25 


150 


25 


21.5 X 22.5 


1 - 8 


145 


3.50 


200 


25 


23.3 X 24.5 


1 - 9 


135 


5.50 


250 


26 


24.5 X 25.3 


1 - 10 


135 


8.00 


300 


26 


26.5 X 27.0 


1 - 10 


145 


11.00 


400 


27.5 


27.5 X 24.5 


1 - 12 


115 


14.00 


500 


27.5 


29 X 30 


1 - 12 


145 


18.00 



425. The Cost Of A Cooling-Tower, erected in place, may 
(Practical Engineer, 1916) be from $6 to $7 per kilowatt 
of the power developed by the plant. Or, otherwise, from 
$4.50 to $5.50 per horse power of the engines to be served. 
These values are based on the assumption of a 26-in. vacuum 
in condenser operation. 



QUESTIONS ON DIVISION 10 

1. Why is recooling of condensing-water desirable? 

2. What phenomena are employed in the recooling of condensing-water? 
3 What factors determine the effectiveness of recooling apparatus? 

4. What is relative humidity? 

5. How is the relative humidity of the air determined in practice? 



362 STEAM POWER PLANT AUXILIARIES [Div. 10 

6. Which is most conducive to the recooling of condensing-water — high or low relative- 
humidity? Why? 

7. What three devices or methods are commonly used for re-cooling condensing-water? 
Under what conditions would each be most advantageous? 

8. Explain the operation of a spray-fountain. 

9. What is the per cent, of water-loss from a spray-fountain, relative to the amount 
of recooling effected? 

10. How may spray-fountains be protected from water-loss by high winds? 

11. What is the usual depth of cooling-ponds? 

12. Can spray-fountains be used where ground space is unavailable? How? Explain. 

13. What per cent, of the total power developed by the plant is required for spray- 
fountain operation? 

14. How may the power required for elevating the condensing-water to an overhead 
spray-fountain be compensated for? 

15. What are the essential principles of cooling-tower operation? 

16. What are the four general classes of cooling-towers? 

17. What average per cent, of efficiency may be obtained in cooling-tower operation? 

18. How does the water-loss from a cooling-tower compare with that from a spray- 
fountain? 

19. What per cent, of the re-cooling in a cooling-tower is generally due to evaporation? 
How is the remaining per cent, of the re-cooiing effected? 

20. In what respect does an atmospheric cooling-tower differ from a natural-draft 
closed cooling-tower? 

21. What advantages result from arranging a cooling-tower so that it may be used 
with either .forced or natural draft? 

PROBLEMS ON DIVISION 10 

1. The air entering a cooling-tower has a dry-bulb temperature of 70 deg. fahr. 
and a wet-bulb temperature of 60 deg fahr. The air leaving the tower has a dry-bulb 
temperature of 90 deg. fahr. and a wet-bulb temperature of 88 deg. fahr. What is 
the relative humidity in each case? What weight of water does the air absorb, per cubic 
foot, while passing through the tower? 

2. The quantity of water circulated through the steam condensers of a 1,000 h.p. 
engine plant is 40 lb. for each pound of steam condensed. The engines consume 15 
lb. of steam per horse power per hour. What should be the area of a simple cooling 
pond to re-cool the condensing water in summer? What should be the area if the pond 
were equipped with a spray-fountain? 

3. In Problem 1 the air re-cools 800 gal. of condensing water per minute through 20 
deg. fahr. The water enters the tower at 105 deg. fahr. and leaves at 85 deg. fahr. 
It is assumed that 20 per cent, of the heat abstraction is due to convection and radia- 
tion, while the remaining 80 per cent, is due to evaporation. What volume of air flows, 
per minute, through the tower? What is the efficiency of the tower? What is the per 
cent, of evaporation-loss? 

4. Assuming that the cooling-tower of Problems 1 and 3 is furnished with a cypress- 
board checker work (Fig. 329), what is the free area through the tower? If the checker 
work is of square cross-section, what are its base-dimensions? 

5. A spray-fountain, fitted with 2-in. nozzles, is to operate under a pressure of 6 lb. 
per sq. in. The quantity of water circulating through the condensers is 40,000,000 
gal. per day of 24 hr. How many nozzles are needed? What pond area is required? 



DIVISION 11 
STEAM-PIPING OF POWER PLANTS 

426. The Steam-Piping Of A Power Plant Generally Com- 
prises Two Separate Systems: (1) The live-steam piping. 
(2) The exhaust-steam piping. The live-steam piping is 
usually designed to convey live steam, either saturated or 
superheated, from boilers to engines and other steam-con- 
suming apparatus at pressures from about 100 to 300 lb. per 
sq. in. It is, therefore, built of the heavier and stronger grades 
of pipe and fittings. The exhaust-steam piping is usually 
designed to carry exhaust-steam, from turbines, reciprocating 
engines and from steam pumps, under pressures ranging from 
less than atmospheric to perhaps 10 lb. per sq. in. It may, 
therefore, be built of comparatively light pipe and fittings. 

427. The Materials For Steam -Piping comprise mainly: 

(1) Wrought iron. (2) Mild steel. (3) Cast-steel. (4) Cast- 
iron. (5) Malleable iron. Wrought-iron pipe is much favored 
on account of its reputation for ductility and durability. 
Pipe made of mild steel produced by the open-hearth process 
is, however, commonly con- 

, , , i-ii K— -556-m. — -H K--5.5G-in.--H k— 5.56-in. -4 

ceded to be equal in all i i i i i i 

respects to wrought-iron jf^T!! \ I*"* 8 * 1 "! J j .i^£j 

pipe. Cast-steel, cast-iron U^"^^^ 1 'J^^^^^ 

and malleable iron are used | I f\ 

mainly in the making of ^^ J/ \& ?> ^Jy 
fittings. ^<^z^ ^ZZZ*' 

Ana Tt c^ j r\£ I- Standard I- Extra Heavy 

428. The Grades Of y 

Steel And WrOUght-IrOn FlG - 33S.— Inside And Outside Diameters Of 
t^. /- N rv, 7 , Three Grades Of Wrought Iron 5-In. Pipe. 

Pipe are: (1) Standard. 

(2) Extra heavy. (3) Double extra heavy (Fig. 338). The 
thickness, and weight per unit of length, of the three grades 
of pipe increase in somewhat irregular ratios. 

Example. — The thickness of a 5-in. pipe (Fig. 338) advances from 
0.247-in. in the standard grade to 0.355-in. in the extra heavy grade, 
and 0.71-in. in the double extra heavy grade. The weight of a 5-in. 

363 




364 



STEAM POWER PLANT AUXILIARIES [Div. 11 



pipe, per foot of length, advances from about 12.5 lb. in the standard 
grade to 17.6 lb. in the extra heavy grade, and 32.5 lb. in the double extra 
heavy grade. Approximately similar ratios are noted throughout the 
tables of sizes. 

Note. — The Sizes Of All Steel And Wrought Pipes, up to 12-in. 
refer to the nominal inside diameters. Above 12-in., the sizes refer to 
the actual outside diameters. These large pipes are made in several 
thicknesses from Y± in. to 1 in. The thinner pipes are used for the lower 
pressures and the thicker for the higher pressures. In purchasing, this 
large pipe is specified by both its outside diameter and thickness. 

429. The Grades Of Pipe Fittings Commonly Used In 
Steam-Piping Systems are: (1) Standard cast-iron. (2) 



B-3 in. Open Return c-Mediwn- D-Eccentric 
A-Y-Bend; Bendj. Sweep Double Tee- 
JJranch Elbows. 
c 




'-Inlet 



\< 6i" H _ -• - ""°'p! ms '.Branch- 
Initial Indicating \ collar- - 
Standard Fitting- 





E-Reducing I 
ged Lateral 



^/-F-Lony-Rcidius .'G-Side Outlet.-'', 
Flcm-' Fihnw '■ Elbow 



Elbow 
Straight Face : - 



Fig. 339. — Standard Cast-iron Fittings. 



E - Street 
Collar-. B-45Deg. D-2'mXlose Elbow. 
A-CrosSv\-,,Elbow.. c-Reducer-. -Return Bend Xn 
-■it,-- ^\ii - V r. 



•Initial Indicating 
Standard Fitting 






Male End 



Female ■ 
End- 



Fig. 



340. — Standard Malleabk 
Fittings. 



Iron 



Standard malleable iron. (3) Extra heavy cast-iron. (4) 
Extra heavy malleable iron. (5) Extra heavy cast-steel. (6) Low- 
pressure cast-iron. Standard cast-iron fittings (Fig. 339) 
are designed for steam pressures up to 125 lb. per sq. in. 



■Raised Face 
rl-Base / MSm ' 
.Elbow ; *., Sweep 



Side Outlet 





-Collar 



Fid. 341. 



-Extra Heavy Cast-iron 
Fittings. 



Fig. 342.- 



-Extra Heavy Malleable Iron 
Fittings. 



Standard malleable iron fittings (Fig. 340) may be used 
for steam pressures up to 150 lb. per sq. in. Extra heavy 
cast-iron fittings (Fig. 341) are intended to withstand steam- 
pressures up to 250 lb. per sq. in. Extra heavy malleable 



Sec. 430] STEAM-PIPING OF POWER PLANTS 



365 



iron fittings (Fig. 342) are safe for steam pressures up to 
250 lb. per sq. in. Extra heavy cast-steel fittings (Fig. 343) 
are safe under a steam-pressure of 350 lb. per sq. in. and a 
total steam-temperature of 800 deg. fahr. Thus they are 
available for use in piping for superheated steam. Low- 

>-l - Reo/wc ing Double 
• Sweep Tee 



■rl -Reducing n- SO Dec/. 
v v, Cross Elbow-. 




5^w> ! Flange- I 
Raiseol Force--' 

Fig. 343. — Extra Heavy 
Cast-Steel Fittings. 



V ' '-Flange f 
. '• Straight Face .*' 

Fig. 344. — Low-Pressure Cast- 
iron Fittings. 



pressure cast-iron fittings (Fig. 344) are suitable for steam 
pressures up to 25 lb. per sq. in. They may be used in 
exhaust-steam systems. Their use in live-steam systems, 
even where the pressure does not exceed 25 lb. per sq in., is 
inadvisable. 

430. The Pipes Commonly Used In Steam-Piping Systems 
Are Classified According To Three Different Types Of Con- 



Circular Die-> 




Skelp- 
Fig. 345.— Method Of Forming Butt- Welded Pipe. 



struction: (1) Butt-welded. (2) Lap-welded. (3) Riveted. 
In the making of butt-welded pipe (A -Fig. 345) the squared 
edges of the skelp, B, are brought to a welding heat. The 
end of the pipe is then formed C- and the edges are pressed 
together D- by drawing the skelp through a circular die; 



366 



STEAM POWER PLANT AUXILIARIES [Div. 11 



M. In the making of lap- welded pipe, the edges of the skelp 
are scarfed (A -Fig. 346) and the skelp is rolled into tubular 
form. The skelp is then brought to a welding heat and is 

(Open Overlapped Edges Welded overlctppedjdges-, 



Scarfed 
Edges-, 




Shank of 
Cast Iron 
Mandrel-, 



>t * — ^^^^^^^ c frf?-V - - ----- - - A _ipr?dn 




Fig. 346.— Method Of Forming Lap-Welded Pipe. 

passed (£-Fig. 346) through a circular groove in the welding 
rolls. The weld is made by squeezing the overlapped 
scarfed edges together between the walls and a cast-iron 




Fig. 347.— Straight-Riveted Steel Pipe. 

mandrel. Riveted pipe (Figs. 347 and 348) is made of sheet 
steel. It may be used for exhaust-steam mains. It should 
not be used in live-steam systems. 




. 



•--Spiroil Seams — - 
Fig. 348. — Spiral-Riveted Steel Pipe. 

Notes. — The Strength Of A Butt- Weld is about 73 per cent, of 
the strength of the plate which it joins. The ultimate strength of a 
butt-weld in a steel pipe is about 41,000 lb. per sq. in. The ultimate 
strength of a butt-weld in a wrought-iron pipe is about 29,000 lb. per 
sq. in. 



Sec. 431] STEAM-PIPING OF POWER PLANTS 367 

The Strength Of A Lap- Weld is about 92 per cent, of that of the 
plate which it joins. The ultimate strength of a lap-weld in a steel pipe 
is about 52,000 lb. per sq. in. The ultimate strength of a lap weld in a 
wrought-iron pipe is about 31.000 lb. per sq. in. Lap-weld pipe may be 
used for all purposes of live-steam piping. It is from 40 to 45 per cent, 
more expensive than butt-weld pipe. 

431. The Trade Meanings Of "Wrought-iron Pipe" And 
"Steel Pipe" are not generally understood. Steel pipe is 
(Power, Dec. 14, 1920, page 948) commonly known and 
billed in the trade as " wrought pipe." Jobbers and contract- 
ors are prone to install steel pipe instead of the more expensive 
wrought-iron material even when the latter is specified. They 
are able to make the case in court on the plea ' ' wrought-iron 
pipe" is a trade term meaning either wrought-iron or steel 
pipe as distinguished from cast-iron pipe. The Executive 
Committee and Advisory Board of the National Pipe and 
Supplies Association, in order to prevent the confusion which 
is heretofore existed, recommends the terminology employed 
by the American Society for Testing Materials: (1) Welded 
wrought-iron pipe. (2) Welded steel pipe. If this standard 
terminology is followed the meanings then are: — (1) That 
welded pipe is pipe which is welded no matter what it is made 
of. (2) That welded steel pipe is pipe made by welding steel. 
(3) That welded wrought-iron pipe is pipe that is made of 
wrought iron by the welding process. (4) That wrought-iron 
pipe is pipe made of wrought iron regardless of the process of 
manufacture. 

432. The Safe Working Pressures For Standard Wrought 
Iron And Steel Pipe from data by Crane Co. are as follows: 
J£ in. to J^ in. butt welded, 900 lb. per sq. in. ; % in. to 1 in. 
butt welded, 750 lb. per sq. in. ; 1 in. to 3 in. butt welded, 400 
lb. per sq. in.; 3J-^ in. to 5 in. lap welded, 400 lb. per sq. in.; 
6 in. to 12 in. lap welded, 250 lb. per sq. in. 

Note. — More conservative practice is to limit steam pressures on all 
standard weight pipe to 250 lb. per sq. in. Lap-welded pipe is considered 
somewhat more reliable than butt-welded and is, in general, preferred for 
all steam piping regardless of the pressure. Some engineers specify only 
lap-welded pipe for all steam-power-plant work. 



368 



STEAM POWER PLANT AUXILIARIES [Div. 11 



433. Table Showing Good Practice Regarding Grades Of 
Pipe For Steam-Power Plant Installations. All pipe for pres- 
sures over 125 lb. per sq. in. should be lap-welded. (Con- 
densed from Crane Co. specifications.) 



Pressure, 
lb. per 
sq. in. 
gage 


Service 


Pipe size, 
in inches 


Grade of pipe 


Material 


Up to 125 


Saturated steam. 


Up to 12 in 


Standard Merchant wt. 






14 to 18 in 


$-{ 6 in. thick. 


Steel 




Over 18 in 


% in. thick. 




125-200 


Saturated steam. 


Up to 12 in 


Full standard card wt. 


Steel 




Over 12 in 


% in. thick. 


200-250 


Saturated steam. 


Up to 12 in 


Extra-strong. 


Steel 




Over 12 in 


%6 or % in. thick. 




Steam superheated 
up to 600°F. 


Up to 8 in 


Extra-strong. 


Steel 




Over 8 in 


J^ in. thick. 




Exhaust steam. 


Up to 12 in 


Standard Merchant wt. 


Steel 




14 to 24 in 


At least ^6 in. thick. 



Note. — "Standaed" Pipe (Sec. 428) Is Manufactured In Two Weights: (1) 
Full card weight. (2) Merchant weight. Full-card- weight pipe is manufactured to con-, 
form exactly to the standard dimensions. Merchant-weight pipe is not, strictly, quite 
as thick and strong as is full-card-weight pipe. 

434. Two Principal Types Of Joints Are Commonly Used In 
Steam-Piping: (1) Screwed joints. (2) Flanged joints. Screwed 
joints (A-Fig. 349) between pipe-ends and fittings are usually 
recommended for steam-piping where the pipe-size does not 
exceed 2.5-in. This however depends largely on the pressure 
and service for which the pipe is to be used; generally, for 

pressures below 125 lb. per sq. 
in. the piping connections are 
"screwed" only for pipes up to 
23^ in. nominal diameter. 
Flanged joints (B-Fig. 349) are 
generally easier to manipulate 
than are. screwed joints. They 
afford ready means for disconnecting the various sections of 
a piping-system. Their use is recommended in all steam- 
piping larger than 2.5-in. 

Note. — Flanges commonly form screwed joints with the pipe-ends. 
Hence, the construction of a flanged joint in a pipe-line may and usually 
does entail (B-Fig. 349) the use of one subsidiary screwed joint. 



Screwed p^ 
flam .. Js— 1 
joints ■ •" 



Companion Flanges^ 

Flanged Main 

Joints- ■"■ 




Screwed SubsidiaryJoints- 



Fig. 349. — Screwed And Flanged Joints 
In Steam-Piping. 



Sec. 435] 



STEAM-PIPING OF POWER PLANTS 



369 



■Thread 



.■Chamfer 



435. The Principal Methods of Attaching Flanges To Pipe- 
Ends are : ( 1 ) Threading. ( 2 ) Shrinking. ( 3 ) Flaring or 
lapping. (4) Welding. Threading (J, Fig. 350) consists in 
screwing the flange on the pipe-end. Strength and lightness 
are insured by forcing on the flange until the pipe-end projects 
beyond the flange-face. The pipe-end is then cut off flush with 
the flange-face. In the shrinking method (II, Fig. 350) the 
pipe-end is turned truly cylindrical. The flange is bored to a 
shrink-fit and the face-end of the bore is chamfered. The 
flange is then heated to redness, and is slipped over the pipe- 
end until the end projects beyond the flange-face. When the 
flange has cooled somewhat, the pipe-end is beaded into the 
chamfer with a ball-peen hammer. The pipe-end is finally 
turned off flush with the flange-face. 

In the flaring or lapping 
method (III, Fig. 350) the •<»*** ***L 
flange is bored slightly 
larger than the outside 
diameter of the pipe. The 
end of the pipe is flared 
or belled. An abruptly 
flared end (III, Fig. 350) 
is called a lapped end. 
The flange fits loosely 
around the pipe and forms 
a swivel- joint with the lap. 
This imparts flexibility to 
the structure when the 
flange is bolted tightly to a flanged fitting. One method of 
welding (IV, Fig. 350) consists in heating both the flange and 
pipe-end to a welding heat and squeezing them together 
under heavy pressure, into a single mass. Flanges may also 
be arc-welded or acetylene welded to the pipe-ends. 

Note. — Pipe-end flanges are commonly called companion flanges. 

Note. — The cost of an extra-heavy forged-steel welded flange being re- 
garded as a basis of comparison, or as 100 per cent., the relative costs 
of other types of attachment of extra-heavy flanges, made of different 
materials, may be expressed as follows: 




H-Ftared or Lapped 
Fig. 350 



■JZ-W\<ted 



Methods Of Securing Companion 
Flanges To Pipe Ends. 



24 



370 



STEAM POWER PLANT AUXILIARIES [Div. 11 



Forged-steel shrunk, 120 per cent. Cast-steel shrunk, 105 per cent. 

Cast-iron shrunk, 60 per cent. Forged steel flared or lapped, 95 per cent. 

Cast-steel flared or lapped, 75 per cent. Malleable iron flared or lapped, 

55 per cent. Cast-iron flared 
or lapped, 50 per cent. Forged- 
steel threaded, 95 per cent. 
Cast-steel threaded, 60 per 
cent. Malleable-iron threaded, 
45 per cent. Cast-iron 
threaded, 25 per cent. 

436. Low-Resistance 
To Steam-Flow In The 
Turns Of A Piping System 
Is Facilitated By The Use 
Of Pipe-Bends (Fig. 351). 
Bends (Fig. 352) are also 
used to absorb the con- 
traction and expansion 
movements of piping. The radii of pipe-bends (i?-Fig. 352) 
should always be as great as circumstances will permit. The 




Fig. 351. — Standard Pipe-Bends For Making 
Turns In Piping Systems. 



A- Double Offset- - 
Expansion U-Bend 



T- Tangent on 

Straight Part 

of Pipe- -, 



U-Bend- 




Fig. 



352. — Bends For Taking Up Expansion Stresses In Piping Systems. 

longer the radius, the greater the flexibility at the bend. Also 
the larger the bend the less the liability of buckling the pipe 
when forming the bend. 



Sec. 43^ 



STEAM-PIPING OF POWER PLANTS 



371 



Notes. — The Minimum Advisable Radius For A Pipe-Bend in a 
steam-line (.R-Fig. 352), for pipe sizes from 2.5-in. to 16-in., is five times 
the nominal diameter of the pipe. 

The Minimum Advisable Lengths Of The Tangents Or Straight 
Parts Of Pipe-Bends (Fig. 352) when the companion flanges are either 
threaded (7 Fig. 350) or shrunk (II, Fig. 350) on, increases, in regular 
progression, from 4-in. for 2.5-in. pipe to 11-in. for 9-in. pipe and to 
18-in. for 16-in. pipe. 

"When the flanges are flared or lapped (III, Fig. 350) the tangent- 
lengths range from 6-in. for 2.5-in. pipe to 9-in. for 9-in. pipe and to 
18-in. for 16-in. pipe. When the flanges are welded, the range of tangent- 
length is from 5-in. for 2.5-in. pipe to 6-in. for 9-in. pipe and to 8-in. for 
16-in. pipe. 

437. Three Methods Are Available For Distributing The 
Steam Supplied By A Boiler Plant which consists of more than 



Cross - 




( ^i Compound Engine-^ 

Quarte r ^ , i + ,-'~ '"'.'"" Separator- 
igrc MSllllHIM ; i iMillliilll ^fffe-^ y y 

<t HT^P 3 ^ Quarter * 

Bends-. 



JZM 



SHigh 
Speed 

Engine \ 



Stop Jtseparator- 
Valves^ " 



ti 


f- 






I 






£ 


I 


i 


I 



U-Bends 



I.i I 

# # 

H 



- S top and C heck- 
' Valves 



Fig. 353. — Single Header System Of Steam Piping. 



one boiler unit: (1) The single header (Fig. 353). (2) The 
loop header or duplicate headers (Figs. 354 and 355). (3) The 
unit group (Fig. 356). The single header is the least expensive 
to install. It is, however, the least convenient arrangement. 
Thus, if it were necessary to repair the section of main piping 
between boilers C and D (Fig. 353) boilers A, B and C would 
not be available for supplying the prime movers to the right 
of the defective section, nor would boilers D, E, F, G and H be 



372 



STEAM POWER PLANT AUXILIARIES [Div. 11 



available for supplying the apparatus to the left. . With the 
duplicate headers (Figs. 354 and 355) the cross-connections 
and the arrangement of stop valves insures unrestricted use 
of all the boilers and prime movers, even though it be neces- 
sary to isolate a section of the piping for repairs. 



Blind Stop Valve in 
,-Flanges. - -cross -connection 



., -Double Sweep 
f Tees 



Ordinary 
Stop Vah ' 



Sptflft 




Stop and Check Valves 



Poilers- 



Fig. 354. — Proper Method Of Connecting A Set Of Boilers To Duplicate Main Headers 
Continuation Through Engine-Room Of Headers X and Y Is Shown In Fig. 355. 



/Double-Sweep Tees 



Stop-Valves in 
,.-- Cross- Connections- • 




..--Main Headers 
Single-Sweep Tees—^ 



'•Valves 



Steam Pipe V .Steam Pipes 
to Compound jf* to Turbines 
Engine 




Fig. 355. — Continuation Of Duplicate Headers X and Y, Fig. 354, Through Engine 

Room. 



438. With The Unit-Group Arrangement (Fig. 356) each 
engine is piped directly to an individual set of boilers, usually 
four, as boilers A, B, C and D, or E, F, G and H, or /, J, K 
and L. Equalizer pipes are, however, employed to bond the 
piping of all the boilers in a single system. These equalizers 
are, usually of the same size as the main pipes leading to the 
engines. 

Note. — Pressure-equalization throughout the system is the sole func- 
tion of the equalizer or header-pipes (Fig. 356). They are not designed 



Sec. 439] 



STEAM-PI VISC, OF POWER PLANTS 



373 



to provide storage space, as the headers in the older arrangements 
(Figs. 353 and 354) are, in a measure, required to do. Hence, it is par- 
ticularly advisable, where the unit-group method of piping is used, 
that ample receiver-separators be installed close to the engine throttle- 
valves. 



•Turbines- 



u 



r~ —i 



zz: 



r 



WIHff: 



Double Offset ] 

Expansion Benols—" 






4+ § 



Boilers- 




ffi: 



Gate Valves- 



(rh Division 



■ 



■' ' T ///A 



^=r 



Firing-Aisle 




Automatic Sfop---- 
-.-and Check Valves"' 




■Boilers 



Fig. 356. — Unit System Of Main Steam Piping Showing Three Unit Groups. 

439. Steam-Pipe Sizes May Be Determined Graphically 

by means of a chart (Fig. 357) which was devised by 
H. V. Carpenter. 

Example. — Find, graphically, the pipe-size required to supply 30,000 
lb. of steam per hour to an engine if the allowable pressure drop be- 
tween engine and boiler is 3 lb. per sq. in. The boiler supplies the 
engine through a pipe which is 150 ft. long. The operating steam- 
pressure is 185 lb. per sq. in. gage. 



374 



STEAM POWER PLANT AUXILIARIES [Drv. 11 



Solution. — The given rated steam-flow of 30,000 lb. per hr. may, 
allowing a 50 per cent, excess rating, be reduced to (30,000 -f- 60) X 
1.50 = 750 lb. per min. The given gage pressure, 185 lb., is equivalent 
to (185 + 15 =) 200 lb. absolute pressure. Also, the given pressure- 
drop of 3 lb. in 150 ft. corresponds to 2 lb. in 100 ft. From A, corres- 



atiji adjd uj •a.msQ2J c j 2»4M|osqv aloio-t'SAV 




£>0 O 00 — ^ K)o 

o o • • — 

4-aaj loaaiouYiH -i2>d Moui zuionbg jza sjounod ui scon sunssaja 



. O a 

" « r 



ponding to 750 lb. on the base line (Fig. 357) proceed vertically upward 
to B on the line of 200 lb. absolute pressure. Proceed thence downward, 
parallel to the oblique lines, to C on the line of 2 lb. pressure-drop. 
Tracing vertically upward from C, the point of intersection, D, with the 
top line indicates the required pipe size to be about 6.7 in., or practically, 
7 in. 



Sec. 440] STEAM-PIPING OF POWER PLANTS 375 

440. A Simple Formula, For Computing The Pipe Size 
Necessary To Deliver Steam At A Given Rate, which is used 
often in practice is given below. In using this formula, a 
steam-flow velocity, which practice has shown will not induce 
an excessive pressure drop, is assumed. Then, the required 
pipe diameter or area may be obtained by substituting the 
other known values : 

/ W 

(97) di = 13.54^ yt~ (diam. inches) 

\ L)V m 

144W 

(98) Ai = -=r — ■ (area, sq. in.) 

■LJVm 

Wherein di = actual internal diameter of pipe, in inches. 
W = equivalent weight of steam flowing through pipe, in 
pounds per minute. D = density of steam at the given pres- 
sure, in pounds per cubic foot. v m = velocity of flow of 
steam in pipe (see Sec. 441) in feet per minute. Ai = in- 
ternal area of pipe, in square inches. 

Note. — The Above Formula May Be Used For Figuring The 
Pipe Size Required For A Reciprocating Engine if the valve cut off 
is known. For example, if 30,000 lb. of steam is used by the engine 
per hour and the cut off is Y±, then the equivalent flow will be appro- 
ximately: 4 X 30,000 = 120,000 lb. per hr. These formulae are not re- 
commended for pipes under 3 in. in diameter. 

Example. — A steam engine which is set for }/% cut off requires 12,000 
lb. of steam per hr. The steam pressure is 125 lb. per sq. in. gage. A 
velocity of 6,500 ft. per min. is allowable in the pipe. What size of pipe 
is required ? Solution. — The steam velocity is equivalent to that when 
the steam flows continuously at the rate of 3 X 12,000/60 = 600 lb. per 
min. Substituting in the above formula : di = 13.54.\/W/Dv m = 13.54 X 
V( 600 ) -f- (0.3107 X 6,500) = 7.3 in. internal diameter or a 7 in. pipe 
is sufficiently large if not too long (Sec. 444). 

Note. — This size pipe will have a maximum pressure drop as computed 
by Fig. 357 of 4 lb. per sq. in. per 100 ft. 

441. The Allowable Steam-Flow Velocities Used In Practice 

are about as follows: For average power-plant installations: 
saturated steam, 6,000 to 8,000 ft. per min. superheated 
steam, 8,000 to 12,000 ft. per min. exhaust steam 4,000 ft. 
per min. In large stations the velocity may be, for superheated 
steam, 14,000 ft. per min. for reciprocating engines and 15,000 
ft. per min. for turbines. In one large eastern turbine station 



376 STEAM POWER PLANT AUXILIARIES [Div. 11 

the velocity is 21,000 ft. per min. In another reciprocating- 
engine installation a velocity of 15,000 ft. per min. is used 
without any apparent adverse effect on economy. 

442. The Drops In Pressure In Steam Mains Allowed In 
Practice range up to perhaps 30 lb. per sq. in. from boiler 
to engine. A total loss in pressure of more than 15 per cent., 
however, is not recommended although the friction in the 
mains causes heat which superheats the steam and does not 
represent actual energy lost. Some engineers recommend 
less than 4 lb. per sq. in. drop in pressure per 100 ft. of pipe. 
From Sec. 445 it will be noted that the loss in pressure due to 
a valve is large compared to that in 100 ft. of straight pipe. 

Note. — Where large receiver-separators are installed close to< engine 
throttle valves, pressure-drops of from 1.5 to 2.5 lb. per 100 ft. of pipe are 
permitted. The corresponding velocity of steam-flow is about 9,000 ft. 
per min. Where a pipe-line is very long, the pressure-drop per 100 ft. 
must, obviously, be kept low in order that a fair percentage of the initial 
steam-pressure may be realized at the place of delivery. 

443. The Average Pressure -Drop In Exhaust-Steam Main 
Piping is, ordinarily, from about 0.2 to 0.4 lb. per 100 ft. of 
pipe where the engines are run non-condensing. It is from 
about 0.2 to 0.4 in. of mercury column per 100 ft. of pipe 
where the engines are run condensing with a vacuum of about 
26 in. 

444. The Size Of A Main Pipe Having A Carrying Capacity 
Equal To The Combined Capacities Of Two Or More Branch 
Pipes May, for the same velocity of steam-flow and other 
conditions, be formed by the following formula: 



(99) di m = Vdn 2 + d i2 2 + dtf + etc. (inches) 

Wherein di m = actual inside diam., in inches, of main pipe; 
day di2, diz, etc. = inside diameters, in inches, of branch pipes. 

Example. — Assuming the same velocity of steam-flow in the main and 
branch piping, what should be the size of a header to supply four branches 
having diameters of 3-in., 3^-in., 5-in., and 6-in., respectively? 

Solution. — By For. (99), d im = 

Vdti 2 + d<2 2 + d<3 2 +&4 2 - V3 2 + 3.5 2 + 5 2 + 6 2 = 9.1-in. 
Hence, a 10-in. pipe is required, since this is the next larger size to the 
value found. 



Sec. 445] 



STEAM-PIPIXG OF POWER PLANTS 



377 



445. The Pressure -Drop Due To The Presence of Globe 
Valves And Right-Angled Fittings In Steam Pipes may be 
taken into account, in the computations for size, by applying 
Briggs formulae, which are as follows: 

di) 
3.6> 



(100) 
(101) 



L v = lUdi ~ (l + 

L , _ 7M< + (i + ™) 



(inches) 
(inches) 



Wherein L v = pipe-length, in inches having resistance equiva- 
lent to that of one globe valve, L e = pipe-length, in inches, 
having resistance equivalent to that of one standard 90 deg. 
elbow. d{ = internal diameter of pipe, in inches. 

Note. — Gate Values And Pipe Bends Produce Pressure Drops 
only equal to the length of pipe they actually contain. That is, a pipe 
bend which is 30 in. long measured along its circular center line would 
produce the same pressure drop as a straight, 30-in. length of the same- 
size pipe; a gate valve measuring 8 in. from face to face would introduce 
the same pressure drop as an 8-in. straight length of pipe of the size into 
which the valve is designed to be fitted. These drops may be read from 
Fig. 357. 

Example. — To how many feet of pipe-length would the resistance 
offered by one globe stop-valve and two standard 90-deg. elbows in a 
7-in. steam-line be equivalent? 

Solution.— By For. (100), L v = lUd t ■*- (1 -f (3.6/d<) = 114 X 7 v 
[1 + (3.6 + 7)] in. = 527.08 = 527.08 
4- 12 = 43.9 ft. By For. (101), L e = 
76d + (1 + 3.6/d») = 76 X7v(l + 
3.6 -f- 7) = 351.39 in. = 351.39 -5-12 = 
29.28 ft. Hence, the total equivalent 
pipe-length = 43.9 + 29.28 X 2 = 
102.46 ft. 



Double. 
Extra- 
Heavy- 
Brass 
Hippies 



Horizontal 

Runs of 

Steel or 

Wrought- 

Iron 

Pipe-.. 




446. Linear Expansion In 
Steam Pipes tends to produce 
bending, buckling, and tensile 
stresses in the piping. Strains 
due to these stresses are ob- 
viated (Figs. 352, 358, 359 and 360) by the use of compensating 
devices. 

Note. — Expansion Slip-Joints (Fig. 359) are mainly used with very 
large pipes, and where space prohibits (Figs. 352 and 358) long-radius 



■Cast- Steel 
Elbows 

Fig. 358. — Double-Swing Or Swivel 
Joint For Taking Up Expansion In Pipe 
Lines. 



378 



STEAM POWER PLANT AUXILIARIES [Div. 11 



bends, or swivel joints. When slip-joints are necessary, binding in the 
joint, due to sagging of the pipe, must be guarded against by erecting 
substantial supports at each end. Also, the pipe must be securely an- 
chored to prevent the steam-pressure from forcing the joint apart. 



■Horizontal Pun of Pipe— 
(Packing Glands-, 




Sleeve- 



,-'_ Anchor Base-' \ '--Sleeve 
'Fibrous Packing-' 



Fig. 359. — Double-Slip Expansion Joint. 



r'Joinf 



(Corrugated Copper 
j Casing ; for Taking up 




'Polished 
. . Steel 

'Screws Lining 



Steel Reinforcing 
Rings * 



Fig. 360. — Corrugated Expansion Joint. 



447. The Linear Expansion Occurring In Steel And Wrought- 
Iron Steam Pipes may, for given lengths of piping and ranges 
of temperatures, be found by the following formula : 

(102) I = eiLTf (inches) 

Wherein: 1= the linear expansion of the pipe, in inches, ei = 
the coefficient of linear expansion (see note below). L = the 
original length of steam pipe, in inches. T/ = the change of 
temperature, in degrees fahrenheit. 

Note. — The coefficient of linear expansion (ei) for charcoal iron is 
0.000,006,86; Bessemer steel, 0.000,006,99; seamless open-hearth steel, 
0.000,006,88; cast iron, 0.000,006,2; cast steel, 0.000,006. 

Example. — What will be the linear expansion in a straight 150 ft. 
line of Bessemer steel pipe when steam at a pressure of 125 lb. per sq. 
in., gage, is admitted, if the pipe has a temperature of 60 deg. fahr. at 
the time of erection? Solution. — A table (see the author's Practical 
Heat) of the properties of saturated steam gives the temperature at 
125 lb.per sq. in. gage as 353.1 deg. fahr. By For. (102) I = e t LT f = 
0.000,006,99 X (150 X 12) X (353.1 - 60) = 3.69 in. 

448. The Least Length Of Pipe Necessary For A Bend Or 
Loop To Take Up The Expansion In A Run Of Pipe Of Given 
Length may be found by Rayne's formula, which is as follows : 

(103) L h = 0MSVdoL p T f (feet) 
Wherein: L b = least length, in feet, of pipe required for bend. 
do = external diam., in inches, of pipe. L p = length, in 



Sec. 449 J STEAM-PIPING OF POWER PLANTS 



379 



feet, of pipe-line, 
heit. 



T f — temperature rise in degrees Fahren- 



Example. — What is the least length of pipe that should be used in 
making a double-offset expansion U-bend (A, Fig. 352) to be installed in 
a straight 150-ft. run of 6-in. pipe designed to carry steam at 150 lb. 
pressure, gage, if the temperature of the piping when erected is 60 deg. 
fahr.? 

Solution. — A table of the properties of saturated steam gives the tem- 
perature at 150 lb. pressure, gage, or 165 lb. pressure, absolute, as 366 
deg. fahr. The outside diam. of a 6-in. pipe is 6.625 in. By For. (103) Lb 
= 0.043 VdoL p T f = 0.043 X V6.625 X 150 X (366 - 60) =23.7 ft. 

Note. — The results obtained with the preceding formula can be applied 
directly only with steam pipes of the smaller sizes. With the larger 
sizes, it may be necessary to increase the computed lengths in order to 
conform to the minimum allowable ratio (Sec. 436) of pipe-diameter to 
radius of curvature, and to the prescribed tangent lengths. 

449. Vibration In Steam-Piping is generally caused by a 
pulsating steam-flow. The pulsations may be due to the 



■Steam Mam 




■3 in Extra-Heavy Pipes for Absorbing . -' 
Lateral Vibration of Steam Main-- ' ' 

Fig. 361. — Devices To Prevent 
Transmission Of Pipe- Vibration. A, 
Floor-Support For Use Where Space 
Is Ample. B, Double-Spring Hanger 
For Use Where Head-Room Is Limited. 




■Pips Flange Embedded 
In Concrete Floor 

Fig. 362. — Devices To Prevent Trans- 
mission Of Pipe- Vibration. A, Floor-Support 
For Use Where Space Is Restricted. B, Simple 
Spring Hanger With Safety Device. 



alternate opening and closing of the admission valves of reci- 
procating engines. Transmission of the vibration to the 
foundations and walls of buildings may be prevented (Figs. 
361 and 362) by special supporting devices. 



380 



STEAM POWER PLANT AUXILIARIES [Div. 11 



450. Various Devices Are Used For Staying And Supporting 
Steam-Piping in order to prevent deflection and vibration. 
These devices mainly comprise: (1) Plain hangers (Fig. 363). 



-I- Beam 




Fig. 



363. — An Ordinary 
Pipe Hanger. 



i -Binding- 
u Roots* 




Bracket- 



Fig. 364.— Wall-Bracket, 
With Binding Rolls, For 
Supporting Steam Main. 




Fig. 365.— Simple Floor- 
Stand For Supporting 
Steam Main. 



(2) Wall-brackets (Fig. 364). (3) Floor stands (Fig. 365). 
(4) Anchors (Fig. 366). (5) Counter-balancing hangers (Fig. 
367). Plain hangers should be free to swing (Fig. 363) in the 



-Flat Iron 
Anchor-Band 




Guides (Secured to Wall) - 

■Counterweights- 
Steel Levers-. 




Fig. 366. — An Ordinary Pipe- 
Anchorage. 



Fig. 367. — Method Of Suspending And Counter- 
balancing Expansion Loops In Steam Mains. 



direction of the length of the pipe. Also, they should also be 
provided with a means for height-adjustment. Wall-brackets 
with roll-binders (Fig. 364) allow for free linear expansion of 
the pipe, but prevent lateral movement. Such binders should 



Sec. 451] 



STEAM-PIPING OF POWER PLANTS 



381 



be used in supporting the ends of horizontally-placed long- 
radius bends. An anchor (Fig. 366) is designed to hold the 
pipe immovable, at the place of anchorage, against expansion 
stresses. Counter-balancing hangers (Fig. 367) are designed 
to sustain the weight of expansion-loops, while giving free 
play to the rise and fall of the loops under alternate expansion 
and contraction. 

451. The Heat Losses From Bare And Insulated Steam 
Pipe are as follows (based on Marks' Mechanical Engineers' 
Handbook) : 

Insulation No. 1 is of a hard fire-proof variety of asbestos of relatively 
poor insulating value. No. 2 is sponge-felted asbestos. The conductivity 
of most insulation for pipes is intermediate between these two sets of 
values. The insulation is assumed to be about 1 in. thick 



Temperature difference, pipe and 
air, deg. fahr. 



50 



100 



200 



300 



400 



500 



Loss in B.t.u. per hr. per 
deg. fahr. temperature 
difference per sq. ft. of 
pipe surface. 



Bare 
pipe 



Insul- 
ation 
No. 1 



Insul- 
ation 
No. 2 



1.95 



0.63 



0.34 



2.15 



0.65 



0.35 



2.665 



0.715 



0.369 



3.26 



0.781 



0.391 



4.035 



5.18 



0.856 0.967 



0.414 



0.439 



452. The Condensation Due To Loss Of Heat From Bare 
Steam Pipes may be found by the following formula: 

2.7A,(T fs -T fa ) 



(104) 



W c = 



H, 



(lb. per hr.) 



Wherein: W c = weight of condensation, in pounds per hour. 
Af = area of external surface of pipe, in square feet. Tf< = 
steam temperature at given pressure, in degrees fahren- 
heit. T fa = temperature of surrounding air, in degrees 
Fahrenheit. H v = latent heat of steam at given pressure, 
in British thermal units per pound. 



382 



STEAM POWER PLANT AUXILIARIES [Div. 11 



Example. — The external-surface area of 4-in. pipe is 1.178 sq. ft. per 
ft. of length. What will be the quantity of condensation in 40 ft. of 
bare 4-in. pipe carrying steam at 105 lb. pressure, gage, when the sur- 
rounding air-temperature is 60 deg. fahr.? 

Solution. — A table of the properties of saturated steam (Author's 
Practical Heat) gives the temperature of steam at the given pressure 
as 341 deg. fahr., and the latent heat as 877.2 B.t.u. By For. (104), 
W c = 2.7 A f (T /a - Tfo) -r H v = 2.7 X 1.178 X 40 X (341 - 60) + 
877.2 = 40.75 lb. per hr. 




Fig. 368. — The Holly Steam-Loop For Draining High-Pressure Piping. 

453. Excessive Loss Of Heat From Steam Pipes May Be 
Prevented by covering the pipes with heat-insulating material. 
Incombustible mineral substances, as magnesia and asbestos, 
are commonly used for this purpose. All steam-pipe cover- 
ings should be at least l-in« thick, The heat-loss, with a good 



Sec. 454] STEAM-PIPING OF POWER PLANTS 383 

covering, maj^ be reduced to about 15 per cent, of that occur- 
ring with bare pipe, or even less. 

454. The Condensation In High-Pressure Steam-Piping 
May Be Returned To The Boilers With A Holly Loop (Fig. 
368). The condensation gravitates to a receiver, A, wherein 
it is broken into a spray by passing through a perforated 
plate. Connections should be made to the receiver from all 
parts of the piping system wherein water might become 
pocketed. Due to the discharge of steam from the discharge- 
chamber C, through the vent-pipe, P, and reducing-valve, 
into the feed-water heater, the pressure in the discharge- 
chamber is less than that in the receiver. Hence, a current 
of water-spray, mixed with steam-vapor, ascends through the 
riser R. The steam and water separate in the discharge- 
chamber. The water gravitates to the boilers through the 
drop-leg D. The discharge-chamber is placed at an elevation 
that will insure a sufficient hydrostatic head to overcome the 
excess of boiler steam-pressure over the discharge-chamber 
steam-pressure. Circulation in the loop is started by opening 
valve S. When steam appears, valve S is closed and the 
reducing valve is opened. 

QUESTIONS ON DIVISION 11 

1. What pressures are commonly carried in live-steam piping? In exhaust-steam 
piping? 

2. What are the ordinary materials of steam-piping? 

3. Enumerate the regular grades of steel and wrought-iron pipe. 

4. To what dimensions do the nominal sizes of piping refer? 

5. Enumerate the grades of pipe fittings commonly used. 

6. What is the maximum advisable pressure for malleable-iron fittings? For standard 
cast-iron fittings? For extra heavy cast-steel fittings? For low-pressure cast-iron 
fittings? In extra heavy cast-iron fittings? 

7. How is a lap-weld made in steel or iron pipe? A butt-weld? 

8. For what purpose in power plant steam-piping may riveted pipe be used? 

9. What per cent, of the plate-strength is secured with a lap- weld? With a butt- 
weld? 

10. What is the ultimate strength of a butt-weld in a steel pipe? In a wrought-iron 
pipe? 

11. What is the ultimate strength of a lap-weld in a steel pipe? In a wrought-iron 
pipe? 

12. What is a companion-flange? 

13. How is a companion-flange shrunk on a pipe-end? How welded on? How is the 
pipe-end finished when the flange is threaded on? What kind of a fit does the flange 
make with a flared or lapped pipe-end? 

14. What are the purposes of pipe-bends? What is the minimum advisable radius for 
a pipe-bend? The minimum advisable tangent-length for a 9-in pipe bend with shrunk 
flanges? 



384 STEAM POWER PLANT AUXILIARIES [Div. 11 

15. Which is a tangent-length in a pipe-bend? 

16. Enumerate the principal methods of distributing the steam-output of a set of 
boilers. 

17. What advantage is secured with duplicate main headers? 

18. What are the main features of the unit-group system of steam distribution? 
Why are receiver-separators particularly necessary in the branch pipes to engines where 
this system is used? 

19. What is the commonly-assumed rate of steam-flow for live-steam piping? For 
exhaust-steam piping? 

20. What is the commonly-assumed range of pressure-drop for live-steam piping? 
For exhaust-steam piping? 

21. Describe a slip expansion-joint. A swivel expansion-joint. A corrugated 
expansio n-j oint. 

22. How may transmission of pipe-vibration be prevented? Describe a method of 
support and of suspension to localize pipe-vibration. 

23. Enumerate the common methods of staying and supporting steam-piping. Enum- 
erate the special adaptations of each. 

24. What is the average percentage of heat-saving effected with pipe coverings? 

25. Explain the operation of the Holly steam-loop. 

PROBLEMS ON DIVISION 11 

1. The required maximum steam-output of a boiler is 30,000 lb. per hr. at 150 lb 
pressure, gage. The total length of pipe in the lead to the main header being 40 ft., 
what should be the pipe-size? 

2. Assuming a uniform velocity of flow in the main and branches, what 
should be the size of a main to supply four branches of sizes 2.5-in., 4-in., 5-in., and 
7-in., respectively? 

3. A 6-in. run of steam-pipe contains two globe-valves and one standard 90-deg. 
elbow. What length of 6-in. pipe would offer equivalent resistance to the steam-current? 

4. What minimum length of pipe' is permissible in making an expansion U-bend to 
be used in an 8-in. steam-line, 150 ft. long, carrying steam at 135 lb. pressure per sq. in., 
gage? The temperature of the piping, when erected, is assumed to be 60 deg. fahr. 

6. What will be the quantity of condensation in 30 ft. of bare 10-in. steam-pipe in an 
atmospheric temperature of 90-deg. fahr., if the steam-pressure is 125 lb. per sq. in., 
gage? The external surface area of 10-in. pipe is 2.816 sq. ft. per ft. of length. 



DIVISION 12 



LIVE-STEAM AND EXHAUST -STEAM SEPARATORS 



455. A Live-Steam Separator (Fig. 369) is a device for re- 
moving entrained water from the steam which is conveyed, 
through pipe-lines, from boilers to various steam-consuming 
apparatus, as reciprocating engines and turbines. 



Pipeline 
from Boiler- 



a 



Exhaust- 
Head--' 



Live-Steam 
Separator^ 



Back-Pressure 
Valve — 




■ ^— -T- p-—r-Vj? : 



To Feed-Water 
Heater-. 




Glass ] i 
Water-Cage 



Exhaust- 
Head 
-t -Drain 



■Exhaust 
Pipe 



Exhaust-Pipe 
"—-Drain 



-k— Separator 
Drain 



/■Floor 
r 



<J ' ' ' c £ ' ":■' ■& A a ' ; : ■'••4 <J: 
-a*::<s ;"\-/>;: : :. . -4-. -4.:., . . j. r 



Fig. 369. — Live-Steam And Exhaust-Steam Separators Installed In Engine Piping. 



Note. — Ordinarily, The Steam Issuing From A Boiler Which Is 
Unprovided With Superheating Surface May Contain From 0.3 
Per Cent. To 5 Per Cent. Of Moisture. — If the steam space of the 
boiler is unduly restricted, as where an excessively large number of tubes 
are used in a return-tubular boiler, the percentage entrainment may 
exceed greatly the maximum figure noted above. Similarly, if a properly- 
25 385 



386 STEAM POWER PLANT AUXILIARIES [Div. 11 

proportioned boiler is forced much beyond its rated capacity, the entrap- 
ment may become dangerously excessive. 

Note. — Moisture May Be Carried From A Boiler Either As 
Finely Divided Spray Or As Concentrated Bulks Of Water. It 
may also be due, wholly or in part, to condensation in the pipe-line. 
The quantity so produced will depend largely upon the length of the 
pipe and the effectiveness of the covering. Water resulting from con- 
densation may accumulate in pockets in the piping, whence it may be 
picked up in bulk by the onrushing current of steam. Similarly quantities 
of water in bulk, or slugs of water, may be projected from the boiler by 
the violent priming that may result from a suddenly applied overload, 
or from carrying the water too high in the boiler. 

456. The Purposes Of Live-Steam Separation are: (1) To 
conserve the energy of the steam. (2) To prevent wrecking of 
engines by slugs of water which might be present with the steam- 
supply. (3) To prevent impairment of engine-lubrication by 
wet steam. (4) To protect the valves, pistons and cylinders of 
reciprocating-engines, and the blades and buckets of turbine, 
from the erosive action of wet steam. 

Note. — Moisture Diminishes The Net Thermal Value Of The 
Steam Which Is Delivered To An Engine, and, therefore, the thermal 
efficiency of the engine. It does this by adding to the initial condensation 
in the cylinder and by absorbing whatever superheat may be available 
from expansion. A discussion of this subject is contained in the Author's 
Practical Heat. 

Note. — Admission Of An Otherwise Trifling Bulk Of Water 
To An Engine Cylinder Is Extremely Dangerous if the engine is 
running at high speed. This is due both to the very restricted clearance 
spaces which considerations of economy demand for high-speed recipro- 
cating engines and to the fact that water is practically incompressible. 

Note. — Steam Turbines May Be Seriously Damaged By Slugs 
Of Water Entering With The Steam. The blades and buckets of 
turbines are liable to be stripped by smaller masses of water than such 
as might be required to wreck the cylinders of reciprocating engines. 

Note. — Thorough Lubrication Of An Engine-Cylinder Is Prac- 
tically Impossible When Excessively Wet Steam Is Used. The 
water will gather on the rubbing surfaces and thus exclude the oil. 
Otherwise it will precipitate the oil and flush it out before it can reach 
the rubbing surfaces. 

457. The Economy Of Live-Steam Separation is, aside from 
the considerations previously noted (Sec. 456), mainly a 
question of fuel saving which results from delivering dry steam 
to the prime mover. The loss from initial condensation, due 



Sec. 458] 



STEAM SEPARATORS 



387 



to the effect of wet steam in the engine cylinders, may be 
regarded as approximately 1 per cent, for each 1 per cent, of 
moisture in the steam (Direct Separator Company, Steam and 
Oil Separators). It may also be assumed for turbines that 
for each 1 per cent, of moisture in the steam supplied there is 
an increase of about 2 per cent, in the water rate (Harrison 
Safety Boiler Works, Separators). 

Note. — The Loss Of Efficiency Due To Wet Steam In Turbine 
Operation may be ascribed to the extra friction which the moisture 
creates within the turbine. The added friction apparently necessitates 
supplying an extra pound of steam for each pound of moisture in order 
to maintain a proper velocity of flow. 

Example. — Assuming that 10 tons of coal, at 3 dollars per ton, are 
consumed per day in firing a power plant, the saving which might be 
effected by a 2 per cent, reduction in the moisture content of the steam 
delivered to the engine would annually amount to 10 X 3 X 365 X 
0.02 = $219. 

458. The Principal Operative And Structural Requisites 
Of A Live -Steam Separator in the supply-line to an engine 
are: (1) It should afford the max- 
imum attainable effectiveness of 
separation. The separation 
should be (Table 474), prac- 
tically, 100 per cent, effective 
when the moisture entrained 
with the steam is less than 5 
per cent. It should be at least 
98 per cent, effective when the 
entrainment amounts to about 
20 per cent. (2) Its tendency 
to reduce the pressure of the steam 
should be practically inappre- 
ciable. (3) It should have storage 
capacity equal to about four 
times the volume of the engine 
cylinder. (4) It shoidd be of 
simple and durable construction. 

459. A Live Steam-Separator Is Called A Receiver-Separ- 
ator When It Is Provided With A Relatively-Large Well 
(Fig. 370) . The well serves, both as a receptacle for the water 




■Dram 
I-Hdlf End Section 



I-Siote Section 



Fig. 370. — Vertical Sections Through 
Cochrane Horizontal Receiver-Sep- 
arator. 



388 STEAM POWER PLANT AUXILIARIES [Div. 12 

which is extracted from the steam, and as a reservoir wherein 
an ample volume of steam (Sec. 458) may be continuously 
maintained while the engine is running. 

460. The Steam-Storage Capacity Afforded By A Receiver- 
Separator is of three-fold importance: (1) It operates to 
prevent the vibration to which a long, sinuous, high-pressure 
steam-line might, otherwise, be liable. A prevalent cause of 
vibration of steam-supply lines to engines is the reaction which 
results from the sudden arrest, at cut-off, of the steam-current, 
and the consequent impact of the steam with the back of the 
valve. The constant volume of steam, which a receiver- 
separator may maintain in close proximity to an engine cy- 
linder, acts as a buffer to absorb the shock of such reaction. 
(2) It tends to prevent a drop of pressure between the boiler and 
the engine. The pressure-drop in a steam-supply line may, in 
the absence of storage space close to the engine cylinder, 
amount to 10 per cent, of the boiler pressure. (3) It acts to 
prevent the excessive priming which might, otherwise, attend a 
suddenly applied overload. 

461. An Exhaust-Steam Separator (Fig. 369) is a device 
for removing oil from the steam which has been used in engine- 
cylinders and expelled therefrom. 

Note. — Exhaust-steam separators are commonly called oil-separators 
and oil-eliminators. 

462. The Main Purposes Of Exhaust-Steam Separation 

are: (1) To render the steam suitable for use in open feed-water 
heaters and thereby conserve the heat therein. Oil in the 
feed-water is dangerous to the integrity of steam-boilers. 
(2) To preserve the radiating-effectiveness of exhaust-steam 
heating systems. (3) To preserve the condensing-effectiveness 
of surface condensers. A film of oil in the radiators of a heating 
system, or in the tubes of a condenser, greatly retards the 
transmission of heat from the steam to the external air in the 
one case, or to the cooling water in the other. (4) To render 
available, for boiler feed-water, the discharge from surface 
condensers. 

463. The Economy Of Exhaust-Steam Separation is, aside 
from the principal considerations previously enumerated 



Sec. 464] STEAM SEPARATORS 389 

(Sec. 462), largely a question of the saving which purification 
of the exhaust-steam effects in the cost of water for operating 
the plant. Where the boiler feed-water is taken, without 
cost, from streams or other nearby sources, conservation of the 
water supply is of little moment. But where the boiler-water 
is taken from city mains, the expense of wasting the exhaust- 
water may assume serious proportions. 

Example. — Allowing 14 pounds of feed water per hour per horsepower 
developed by a set of condensing engines the annual water-consumption 
•for this purpose would be about 14 lb. X 24 hr. X 365 days ■*■ 62.5 lb. 
per cu. ft. = 1,962 cu. ft. per h.p. If the water costs $0.50 per 1,000 
cu. ft., and the plant develops a daily average of 10,000 h.p., the annual 
expense for boiler-feed, if the discharge from the condensers were wasted, 
would, therefore, be (1,962 X 10,000 -r- 1,000) X 0.50 = $9810.00. As- 
suming, in this case, that 80 per cent, of the condensed exhaust steam 
were returned to the boilers as clean feed-water, the annual saving would 
be 9,810 X 0.8 = $7848.00. 

464. The Physical Phenomena Involved In The Operation 
Of Steam-Separators are: (1) Expansion. (2) Momentum. 
(3) Elasticity. (4) Capillary entrainment. (5) Absorption. 
These principles, as explained hereinafter, are variously 
applied. The first four are observable in the operation of all 
separators. 

Note. — Expansion, As A Principle Of Separation, is prominent in 
the workings of all receiver-separators. The current of steam expands 
somewhat after issuing from the contracted pipe passage (P, Fig. 370) 
into the relatively-ample volume of the receiver, R. Its density thus 
momentarily diminishes. Hence, it becomes less effective for supporting 
the suspended moisture. The tendency of the water particles to drop 
out of the steam by their own weight is, therefore, increased. 

465. Steam Separators May Be Classified According To 
Their Principal Modes Of Operation as follows: (1) Reverse- 
current separators. (2) Centrifugal separators. (3) Impact or 
Baffle-plate separators. (4) Mesh separators. (5) Gridiron 
separators. (6) Absorption separators. 

466. The Main Operating Principle Of Re verse -Current 
Separators (Figs. 371, 372 and 373) is the momentum which a 
body acquires through propulsion by a force acting along an 
approximately straight line. After entering the separator, 



390 



STEAM POWER PLANT AUXILIARIES [Div. 12 



the moisture-laden current of steam traverses a short distance 
(Fig. 371) in a direct line. Its course is then reversed abruptly. 
The steam readily adjusts itself to the altered direction of 
flow. But the water particles being of much greater specific 
gravity than the steam, are propelled by their own momentum 
to the bottom of the separating chamber. 

Note. — With the separation shown in Fig. 371, removal of the mois- 
ture depends solely upon the whip-snap action which accompanies the 
current-reversal. With the apparatus shown in Fig. 372, two horizontal 
baffle-plates or wings, one projecting laterally from each side of the 



Met- 



Conical Hood* 



Wafer Troughs Surrounding 
Jn/etAnd Outlet Ports*. 
•Outlet Port 
Diaphragm 




Direction Of > 
Steam Current' 
Drainage Ducts Leading 
From Troughs To Bottom 
Of Separating Chamber 

Fig. 371.— Happes Re- 
verse-Current Horizontal 
Exhaust-Steam Separator. 



.' Drain- 
•Receiver 



'-Welded Joint 



Fig. 372. — Welderon Re- 
verse-Current Horizontal 
Receiver-Separator. 



Outlet- 'Outlet Tube] 
Hooded Guard to Prevent-'' 
Upward Creepage of Water 

Fig. 373. — Austin Re- 
verse-Current Vertical 
Live-Steam Separator. 



diaphragmed steam-duct, aid in the separation. With the apparatus 
shown in Fig. 373, the separation is partially effected by impact of the 
current with the hoods. 

467. The Main Operating Principle Of Centrifugal Sepa- 
rators (Figs. 374, 375, 376) is the tangential momentum which 
a body acquires through the action of centrifugal force. The 
steam-current assumes a spiral or twisting motion at the 
instant of its entrance to the separator. The centrifugal 
force thereby developed in the particles of oil or water impels 
them to fly tangentially from the steam-current. Thus, the 
oil or water is flung against the inner surface of the external 
shell, down which it trickles to the drainage outlet. 



Sec. 467] 



STEAM SEPARATORS 



391 



Note. — The device for imparting a twisting motion to the steam in a 
centrifugal separator may be a helix in the throat of the inlet orifice 



Stzam Tnizt- 




Glass 
Water- 
Gage 



-Helix for Throwing. Slof with Over- 
Oil or Wafer 'flapping Eages 
•Inlet ,-•-'"•:••. 




I -End Section 



■Drain- 

II -Side Section 



Fig. 374. — Swartwout Centrifugal 
Steam Separator. 



Fig. 375. — Masher Centrifugal Horizontal Steam 
Separator. 



.-Outlet 




primary Separation by !n> 
[pact of Current with Baffle 

h,et ^ .Upper 




Fig. 376. — Stratton Centrifugal Hori- 
zontal Live-Steam Separator. 



Dram-. 



'■Secondary Separation by 
Reversal of Current 

Fig. 377. — Austin Baffle-Plate Angle 
Live-Steam Separator. 



(Fig. 374), a helix which traverses the interior of the separator from the 
inlet to the outlet (H, Fig. 375), or a spiral web (W, Fig. 376) which 
winds about a central outlet tube. 



392 



STEAM POWER PLANT AUXILIARIES [Drv. 12 



468. The Main Operating Principle Of Impact Or Baffle- 
Plate Separators (Fig. 370, 377, 378, 379, 380 and 381) is the 




Grids for 
intercepting 
Separated Watvr 



Drain 



'Secondary Separa 
tion by Reversal of 
Steam-Current--- 



Fig. 378.—" Austin " Baffle-Plate Under- 
slot Horizontal Live-Steam Separator. 



[Downward Trickle 

; of Water or Oil . . • • -Ribbed Baffles- 

\ between Ribs y -Steam Outlet > 




■ -Drain 



I-Perspectiye Showing 



Fig. 379— "Baum" Baffle-Plate Hori- 
zontal-Steam Separator. 



:Cdd-Water Sprays 



Cold-Water Spray 
Pipe 



.-Flange-Gutter 

for Catching Oil 

Which Gathers in 

Exhaust Pipe 




Connection to 
''•■Auxiliari/ Vacuum Pump 

Fig. 380. — Austin Baffle-Plate Horizon- 
tal Exhaust-Steam Separator For Vacuum 
Service. 



Ribbed Baffle Formed by 
Wall of Cold Water Circulating 
Chamber----. ... , 




Equalizing-' 
Pipe Connection 



Fig. 381.— Baum Baffle-Plate 
Horizontal Exhaust-Steam Sep- 
arator For Vacuum Service. 



elasticity of steam. The entering steam-current (Fig. 377) 
impinges upon the upper baffle, B. Due to its great elasticity, 
the steam rebounds therefrom. But the quite inelastic water 



Sec. 469] 



STEAM SEPARATORS 



393 



Primary Separation 
by Capillar y Action 



adheres to the plate and trickles by capillary entrainment into 
the trough at its lower edge. Thence it flows to the drainage 
outlet, 0. The separation thus far is, however, only partial. 
When the steam rebounds the upper baffle, it strikes the outer 
shell, S. It then rebounds downward, toward the opening to 
the lower baffle, and reverses its direction of flow. Additional 
moisture is thus whipped out by its own momentum. 

469. Corrugated And Fluted or Ribbed Surfaces In Steam 
Separators (Figs. 370, 378, 379, 380 and 381) perform a two- 
fold function: (1) They prevent the sweep of the steam-current 
from scouring the adhering particles of oil or moisture from the 
surfaces. (2) They facilitate the tendency of the separated oil 
or water to trickle downward in a multitude of small individual 
streams. 

470. The Main Operating Principle Of Mesh Separators 
(Fig. 382) is the tendency of fluid particles to entrain and 
form into minute rivulets 
by capillary attraction. 
The entering steam-cur- 
rent, E, impinges directly 
upon the sieve, S, which 
covers the conical top of 
the hood, H, surrounding 
the upper orifice of the 
outlet tube, 0. A portion 
of the water or oil will ad- 
here to the sieve, and, by 
capillary entrainment, will 
pass through its meshes 
to the top surface of the 
hood . The water or oil thus 
deposited flows through the 
drainage tubes D, to the 
collecting-chamber, C. 

Impact with the conical surface changes the form of the 
steam-current to that of an annular sheet which sweeps down- 
ward in the space between the cylindrical wall of the hood, H, 
and the cylindrical sieve or trapping-sheet, T. Nearly all of 
the remaining moisture, or oil, is caught in the meshes (Fig. 




-Conical Top of Hood 
-.-Sieve 
Centering Wings 

Cylindrical Wat! 
of Hood--. 

■Trapping Sheet 

Secondary Sep- 
aration by Rever- 
sal of Current 



Annular Space 
between Trapping- 
sheet and Exter- 
nal Shell ■■ 



Drain 



Outlet- 



Fig. 382.- 



Giass-water-. 
..-<?age 



Outlet 
-Tube 



-Sweet Mesh Vertical Steam Sep- 
arator. 



394 



STEAM POWER PLANT AUXILIARIES [Div. 12 



383) of the trapping-sheet, T. It is thereby entrained in tiny 
streams which flow to the annular space, A, between the trap- 
ping sheet and the shell. Thence it trickles downward to the 



>Sheet>Steel 




Fig. 383.— Sectional Detail Of 
Sweet Steam Separator. 



Side Views of Hol- 
low Grid-Columns- - 



Outlet-. 



Vertical Sections 
of Hollow Grief- 
Columns-. 



in 
Faces of Grid- 
Columns-., 



Drilled- 
Ports 




Oil or Water Dropping' 
from.Chomnels in 
Grid- Columns 



Fig. 384. — Bundy Gridiron Horizontal Steam 
Separator. 



Drilled Ports Leading 
to Hollow Inferiors 
of Columns-^ 



collecting chamber, C. Practically all of the moisture, or oil, 
which still remains in the steam-current will be whipped out 
as the current reverses its direction of flow in passing upward 
to the outlet-tube orifice, 0. The per- 
forated diaphragm, P, prevents the steam- 
current from picking the water, or oil and 
water, out of the chamber beneath. 

471. The Main Operating Principle Of 
Gridiron Separators (Fig. 384) is capillary 
attraction. A series of gridiron separat- 
ing-plates (Fig. 385) is arranged in stag- 
gered formation (Fig. 386) in the path of 
the steam-current. The columns of these 
plates are hollow. Vertical series of 
small cups, or recesses, are cast in the 
faces of the columns against which the 
entering steam impinges. A small hole is drilled from each 
cup to the hollow interior of the column. The particles of 



tnMnlnlMnfl 

U'UUA 

tthtlhl 

mmMm 



Recesses In 

Faces of 

Grid-Columns 1 



Grid- \ 
Columns- 



Fig. 385. — Gridiron 
Separating-P late Of 
Bundy Steam Separator. 



Sec. 472] 



STEAM SEPARATORS 



395 



water, or oil, are projected against the grids and cling thereto. 
The capillary action which then ensues causes them to gather 
in the cups. Thence they trickle through the small ports 
which lead to the channels inside the columns From these 
they fall into the collecting-chamber, C, beneath. 

472. The Operating Principle Of Absorption Separators 
(Fig. 387) depends upon the absorbent properties of certain 
porous or fibrous materials. 



Angle-Plate-^ 



,.-rDri/led Ports, Leading 
lnie> t .''from Recesses toHollow 
'"' \ ! interiors of Grid-Column 




, :R r C ?i?i S f in '■ Vertical Channels 
•■-GridZlumns 'Through Grid-Columns 

Fig. 386. — ■ Staggered Formation Of 
Gridiron, Separating Plates In Bundy 
Steam Separator. 




Exhaust Steal 
Outlet 



|\\\\N\\\\ 

^■Drainage 
Ducts * 



Fig. 387. — Loew Absorption Exhaust 
Steam Separator. 



Note. — Absorption separators are designed only for exhaust-steam 
separation. 

473. The Maximum Efficiency Of Separation Attainable 

(Table 474) with any given type of live-steam separator 
varies according to the quality of the steam as it enters the 
separator. (See Sec. 476 for meaning of efficiency.) 



Note. — The efficiency of a live-steam separator, and, therefore, the 
ultimate effectiveness of separation is benefited by providing an ade- 
quate covering of insulating material. 



396 



STEAM POWER PLANT AUXILIARIES [Drv. 12 



474. Table Showing Efficiencies Obtained In Tests Of 
Live-Steam Separators Of Six Different Makes. 







(From Power, 


May 11, 1909) 






Make 


Steam with less 

than 5 per cent, of 

moisture 


Steam with about 

10 per cent, of 

moisture 


Steam with about 

20 per cent, of 

moisture 




of 
sep- 
arator 


Quality 
of steam 
before 
separa- 
tion 


Qualityi 
of steam 

after 
separa- 
tion 


Quality 
of steam 
before 
separa- 
tion 


Qualityi 
of steam 

after 
separa- 
tion 


Quality 
of steam 
before 
separa- 
tion 


Qualityi 
of steam 

after 
separa- 
tion 


Efficiency, 
per cent. 


A 


97.5 


99.0 


87.0 


98.8 


78.1 


98.8 


60.0 
90.8 
94.5 


B 


9G. 1 


97.4 


90.1 


98.0 


79.5 


98.2 


33.3 
80.0 
91.2 


C 


98.1 


98.5 


89. G 


95.8 


81.7 


- 97.9 


21.1 
59.6 
83.5 


D 


97.7 


97.9 


90.6 


93.7 


78.2 


95.6 


8.7 
33.0 
79.8 


E 


95.6 


95.8 


88.9 


92.1 


82.4 


90.4 


4.5 
28.8 
45.5 


F 


98.0 


98.0 


88.4 


90.2 


79.3 


87.2 


0.0 
15. 5 

38.1 



!Denotes effectiveness of separation. 

475. The Velocity Of The Steam-Current In Transit 
Through A Separator Affects The Efficiency Of The Separator. 

The efficiency diminishes as the velocity increases. If a sepa- 
rator is so designed as to permit an excessive velocity of 
steam-flow through it, its efficiency (Fig. 388) may be practi- 
cally zero. 



Note. — Expansion of the steam (Sec. 460) in transit through the rela- 
tively-large steam space of a separator results in a momentary diminution 



Sec. 476] 



STEAM SEPARATORS 



397 



of the velocity of flow. The initial velocity is, however, restored when 
the steam reenters the outlet pipe if the outlet is the same size as the 
inlet pipe. 



no 





































- Av 
Qu 


i — — j ■ 








































yllty of Steam^ 




























T 












90% 








s^ 


^3 


^* 


























J ! 1 


















^C 






S B 
















res sure of 












F v 




























Steoim= 1 _ 
lOOLb.Gaae 




















E 


















! 






























>• 










D 









































































































"-40 
»J0 
CZO 



£ 1000 2000 3000 4000 5000 6000 1000 8000 

Velocity of Steam- Feet Per Minute 

Fig. 388. — Graph Showing Relation Between Efficiency Of Separation And Velocity O 

Steam Flow. 

476. The Efficiency Of A Live-Steam Separator may be 

computed by the following formula : 



(105) 



100 w„ . 

■& = — yr — (per cent, efficiency) 



Wherein E = per cent, efficiency. W, = weight of separated 




Fig. 389. — Arrangement Of Separator And Appurtenances For Efficiency Test. 

water, in lb. Wz = weight of moisture, in lb., in a definite 
weight of steam delivered to the separator, as determined 
(Fig. 389) by calorimeter and steam-flow tests. 



398 STEAM POWER PLANT AUXILIARIES [Div. 12 

Example. — A steam-flow meter at S (Fig. 389), records a flow of 
16,273 lb. of steam during a certain time-interval. A calorimeter at C 
shows the quality of the steam to be 94.5 per cent. The weight of the 
separated water drawn during the interval from the storage reservoir, R, 
is 530 lb. What is the efficiency of the separator? Solution. — The 
weight of moisture in the steam = 16,273 X (1 - 0.945) = 895 lb. 
Applying For. (105), E = lOOW^/W* = 100 X 530 + 895 = 59.2 per 
cent. 

476A. The Efficiency Of A Live Steam Separator May 
Also Be Computed On The Basis Of The Quality Of The 
Steam Entering And Leaving The Separator by applying the 
following formula : 

(1054) E= 10 ^ 2 ~ Xl) (per cent. efficiency) 

J.UU jC\ 

Wherein: x\ = quality of the steam entering the separator, 
in per cent. x<i = quality of the steam leaving the separator, 
in per cent. 

• 
Example. — In the preceeding example, what is the efficiency of the 
separator if a calorimeter at B (Fig. 389) shows the quality of the steam 
leaving the separator to be 97.8 per cent. ? 

Solution.— By For. (105A): E = 100(x 2 - Zi)/(100 - x{) = 100 X 
(97.8 - 94.5) + (100 - 94.5) = 59.2 per cent. 

477. Exhaust-Steam Separation In A Partial Vacuum May 
Be Facilitated By Wetting The Separating Surface. — This 
may be accomplished in either of two ways: {I) By injecting 
(Fig. 380) a spray of cold water against the surface. (2) By 
circulating cold water within a chamber (Fig. 381) the wall of 
which forms the separating surface. Moisture is thus by 
condensation of a portion of the steam, caused to appear on 
the separating surface. 

Note. — When an engine is exhausting into a partial vacuum the steam 
will have little tendency to condense or to entrain moisture during its 
passage from the engine-cylinder to the condenser.. Hence, all of the 
surfaces which the steam encounters will continue dry. The fine parti- 
cles of cylinder oil will, therefore, due to their very low specific gravity, 
tend to rebound with the steam from the separating surface. But if 
the surface is covered with a film of moisture, the moisture will diffuse 
the oil-particles over the surface and thus cause them to adhere thereto. 



Sec. 478] 



STEAM SEPARATORS 



399 



478. An Exhaust-Head (Figs. 390 and 391) is an exhaust- 
steam separator especially designed for attachment to the 
discharge-end of an engine exhaust-pipe which opens to the 
atmosphere. 



Partial Separation by Impact, 
(Ao/heiyrKe.a nd Capillar y Entrainment 

;Upper Baffle- 
Cone 



Discharge 
■On fee 



Copper 
Drainage 
- -Tubes 




Partial 

Separation by 

Current- Reversal 



Fig. 390.— "Wright" Baffle-Plate Exhaust- 
Head. ' 



Cylindrical Trapping 
firaimae f* e [ Dkchoir g e 
v Hood, i? •' <> Htke 




Fig. 



'Conical 
Trvppto 
Sheet- 
Partial--' ' 
Separation 
by Current ^'--Exhaust-Pipe 
■Reversal Connection 

391.— "Sweet" Mesh Ex- 
haust-Head. 



479. The Purpose Of An Exhaust-Head Is twofold: (1) To 
'prevent pollution of the atmosphere and bejoulment of the roofs 
and walls of buildings by the oil-and-water in the exhaust-steam. 
(2) To muffle the sound of the exhaust. 

480. The Proper Location For A Live -Steam Separator is 
as close to the apparatus which it is designed to serve as the 
piping arrangement will permit. Where the separator is 
used (Fig. 369) in connection with an engine, it should be 
connected directly to the throttle valve. 

481. The Proper Location For An Exhaust-Steam Separator 
depends upon the ultimate disposition of the exhaust. In a 
non-condensing plant, the separator may be installed (Fig. 369) 
in the main exhaust pipe close to the point where it branches 
to the feed-water heater and the radiator heating system. In 
a surface-condensing plant, the separator may be installed 
at any point between the engine and condenser. If a vacuum 
feed-water heater (Sec. 249) is included in the installation, 
and the separator is unprovided with a device for wetting 
the separating surfaces, it may be preferable to place the 



400 



STEAM POWER PLANT AUXILIARIES [Div. 12 



separator between the heater and condenser. The moisture 
which the steam entrains in the heater will thus become 
available for wetting the surfaces. A disadvantage of this 
arrangement is that the heater-tubes will be exposed to 
befoulment by the oil. 

Note. — Exhaust-steam separators are not commonly used in connec- 
tion with condensers in which the steam mingles directly with the con- 
densing water. 

482. The Selection Of A Suitable Live-Steam Separator 

is mainly a question of adapting its shape to structural limi- 
tations. The vertical, horizontal, and angle forms provide 
flexibility of choice in this regard. Otherwise, it is usually 
only necessary, when ordering a separator, to specify the size 
of the steam-pipe, the type of engine and the steam-pressure. 
The proportions adopted by the different manufacturers are 
made conformable to these data. 

Note. — The size of a steam-separator refers to the size of the pipe-line 
in which the separator is installed. 



is 



'From Engine-, 



483. The Selection Of A Suitable Exhaust-Steam Separator 

mainly contingent upon the following information: (1) 

The number and sizes, of the 
engines, including steam-pumps, 
which are to exhaust through 
the separator. (2) The required 
location of the separator (Sec. 
481). (3) Whether the plant is 
operated condensing or non- 
condensing. (4) The pressure of 
the exhaust. (5) The quality and 
quantity of the cylinder oil used. 




Exhaust-' 1 
Pipe 



-A-irr, Exhaust-Pipe 



Fig. 392. — Eclipse Exhaust Steam 
Separator Arranged To Reduce Veloc- 
ity Of Steam Flow. 



Note. — The first and fourth items enumerated above mainly deter- 
mine the velocity of flow through the main exhaust-pipe. The slower 
the steam-flow, the more effective the separation. Adequate separation 
may, therefore, be generally insured by selecting a separator (Fig. 392) 
two or three sizes larger than the exhaust pipe size. 

484. A Live-Steam Separator Should Be Drained Auto- 
matically (Fig. 369) by a reliable steam trap. (See Div. 13.) 



Sec. 48/3 



STEAM SEPARATORS 



401 




Drain Connection-' 



Fig. 393. — Device For Shielding 
Glass Gage From Fluctuations Of 
Steam-Temperature. 



485. Steam-Separators Should Be Equipped With Glass 
Water-Gages (Fig. 369). The glass-gage, G, may be con- 
nected in parallel with a by-pass pipe (P, Fig. 393). The pur- 
pose of this arrangement is to minimize glass breakage; See 
Power 1910. 

Note. — The breakage to which glass gages are peculiarly susceptible 
when attached to steam separators may be due to the frequent and rapid 
changes of temperature to which the glass is subjected. The pressure 
within a separator in the supply pipe of 
an engine may fluctuate through a 
range of perhaps 10 pounds. This will 
be accompanied by a fluctuation in 
temperature which may affect the 
molecular structure of the glass. The 
glass will crystallize quickly and will 
eventually shatter into fragments. 
By locating the gage at a considerable 
distance from the separator and in- 
troducing an intermediary passage 
(P, Fig. 393), sufficient condensation 
may be thereby induced to cause a 
thin film of water to gather on the interior of the glass. This moisture 
will diminish by evaporation as the pressure drops and will augment by 
further condensation as the pressure rises. Thus it may minimize 
temperature fluctuation in the glass. 

486. The Cost Of Steam And Oil Separators: Standard 
horizontal-type oil separators, 2 to 8 in., range in price $8 to 
$36. Vertical-receiver-type oil separators, 2 to 8 in., $13.60 
to $62.00. Standard vertical steam separators, 2 to 8 in., 
$18.40 to $88.00. Standard horizontal steam separators, 
2 to 8 in., $12 to $52. Preceding values (from Mechanical 
And Electrical Cost Data, Gillette and Dana, McGraw- 
Hill) are pre-war costs. During and immediately after the 
great war the prices were advanced from about 100 per cent. 
for the small to 25 per cent, for the large sizes. 

QUESTIONS ON DIVISION 12 

1. What is a live-steam separator? 

2. What percentage of entrained moisture does the steam delivered by a boiler, with- 
out superheating surface, ordinarily contain? 

3. What circumstances of boiler-design and operation principally affect the degree of 
moisture-entrainment ? 

4. What are slugs of water in a steam pipe? What causes the entrained moisture to 
form slugs? 

26 



402 STEAM POWER PLANT AUXILIARIES [Div. 12 

6. What contributory circumstance usually determines the total quantity of moisture 
in the steam delivered to a separator? 

6. Enumerate the chief purposes of live-steam separation. 

7. Through what phenomena, occurring within an engine cylinder, is diminishment 
of the engine's thermal efficiency by wet steam mainly effected? 

8. Why are slugs of water in the steam-supply particularly dangerous to high speed 
reciprocating engines? 

9. In what way may damage occur to a turbine by small masses of water in the steam- 
supply? 

10. How does wet steam affect the internal lubrication of an engine? 

11. What approximate numerical relation exists between the percentage of moisture 
in the steam delivered to an engine and the resulting percentage of loss of economy? 

12. What circumstance apparently explains the loss of thermal efficiency that results 
from supplying wet steam to a turbine? 

13. What are the chief requisites of a live-steam separator? 

14. What is a receiver-separator? 

15. What benefits may attend the use of receiver-separators? 

16. How does a receiver-separator operate to prevent vibration of the steam-supply 
pipe of an engine? 

17. What is an exhaust-steam separator? 

18. What are the principal purposes of exhaust-steam separation? 

19. What is the outstanding consideration with respect to the economy of exhaust- 
steam separation? 

20. What are the physical phenomena which are mainly observable in the operation 
of steam separators? 

21. How does expansion of the steam affect separation? 

22. Enumerate the general classes of steam-separators. 

23. What is the main operating principle of reverse-current separators? Of centri- 
fugal separators? Of baffle-plate separators? Of mesh separators? Of gridiron 
separators? Of absorption separators? 

24. What are the functions of corrugations and ribs on the inner surfaces of steam- 
separators? 

25. What variable factor controls the operating efficiency of a live-steam separator? 

26. What factors determine the operating efficiency of a separator? What factor 
determines the effectiveness of the separation accomplished by a separator? 

27. What effect will diminished velocity have on the efficiency of the separator? 
How may a diminished velocity of flow through a separator be obtained? 

28. Why may advantage result from injecting water into the exhaust-steam separator 
of a condensing engine? 

29. What is an exhaust-head? 

30. What are the functions of an exhaust-head? 

31. What circumstances mainly govern the selection of a proper point of location for 
a steam separator in an exhaust-line? 

32. What considerations are principally involved in the selection of a live-steam 
separator? Of an exhaust-steam separator? 

33. What benefit may result from installing an exhaust-steam separator of larger 
size than the exhaust-pipe size? 

34. How should live-steam separators be drained? 

35. To what inherent circumstance of operation may difficulty of maintaining glass 
water-gages on separators be ascribed? 

PROBLEMS ON DIVISION 12 

1. In a certain locality, coal is available at $4.00 per ton. If 30 tons are normally 
consumed per day, what will be the saving per year if the quality of the steam delivered 
to the reciprocating engines is raised by a separator from 95 to 98 per cent.? 

2. The steam passing to a certain separator has a quality of 93 per cent. If 5,600 lb. 
pass per hour and the separator collects 285 lb. of water, what is the efficiency of the 
separator? 



DIVISION 13 

STEAM TRAPS 

487. Steam Traps are devices for entrapping and auto- 
matically disposing of the water that results: (1) From con- 
densation and entrainage in steam-piping systems (Fig. 394, 395 
and 396). (2) From condensation in steam-heating apparatus, 
(3) From condensation in steam-power apparatus (Figs. 397 
and 398). 



film Eliminator or Live-Steam Separator 
Steamline to Engine-, 



Corrugated Baffle* 

Shields to Prevent 
Steam-Current from 
Picking up Separated 

"Hater- - 

Vent-Pipe- ■- 




Fig. 394. — Nason Bucket-Float Intermittent-Discharge Medium-Pressure Steam Trap 
Installed For Draining A Live-Steam Separator. 



Note. — Steam Traps, In General, May Be Divided Into Two 
Groups: (1) Return traps (Fig. 399) or those which discharge, against 
boiler-pressure, directly into the water spaces of steam boilers. (2) 
Non-return traps (Fig. 400) or those which discharge against normal 
atmospheric pressure, or into receptacles under less than boiler pressure. 

Steam Traps May Be Classified According To The Principles 
Of Operation Chiefly Employed as: (1) Buoyancy traps, which com- 
prise ball-float traps (Fig. 396, 400 and 401) and bucket-float traps (Figs. 
394 and 398). (2) Counierweighied tilting or dumping traps (Fig. 399). 
(3) Expansion traps (Figs. 395 and 397). 

403 



404 



STEAM POWER PLANT AUXILIARIES [Div. 13 



C^b 



■Riser 




Water 

Column 

Strainer-*-'- 



■Alloy Expansion Tube 



Fig. 395. — Kieley Expansion Intermittent-Discharge Steam Trap Draining Radia- 
tion. When T Fills With Water And Cools, It Contracts And Draws In H And P. 
V Is Then Opened By Upward Thrust Of S Against L. When Steam Enters, T Ex- 
pands And Pushes Out H And P. V Is Then Closed By Downward Thrust Of P 
Against L. 




Fig. 396. — Strong Vacuum Trap Installed For Draining Separator In Condensing- 
Engine Exhaust-Line. When F Rises, V Closes And P Opens, Permitting Live Steam 
Or Atmospheric Air Pressure To Discharge Accumulated Water W , 



Sec. 4SS] 



STEAM TRAPS 



405 



Steam Traps May Be Classified According To The Character 
Of Discharge as: (1) Continuous-discharge traps, which are, mainly, of 
the ball-float type. (2) I titer mittcnt-discharge traps. 



Liauid Filled-., 
Votive- Bourdon Tube J 




Closed Feedwater Heater 
Exhaust Pipe from Engine 

Exhaust Steam Inlet 
■ _■*• 



^ *-v 



-Rec/ulottincy Screw 
•Outlet 



*esi 



Exhaust Pipe 

from Boiler ,„.. 

Feeol- mm— 

Pump ,_ 

Trap-.. 



Outlet 





Boiler Feed-\ 
Pump Suction'- 

Exhaust-Steam 
Outlet 



Pass 



Fig. 397. — Marck Expansion Steam Trap Installed For Draining Closed Feed Water 
Heater. When Steam Enters Casing Of Trap, T Expands And Closes V. When Water 
Accumulates In Inlet Pipe, T Contracts and Opens V. Water Enters Casing C Of 
Trap And Passes Out Through O. 

488. The Main Operating Principle Of Return Steam- 
Traps (Fig. 399) is equalization of pressure between the 
interior of the trap and the interior of the boiler into which 

-Drains from Reheating Coils in Exhaust-Steam Receivers--., 
-Drains from Separators in Steam Supply Pipes- ... 




Hmae- 



Fig. 398. — Arrangement Of Tilting-Bucket-Float Intermittent-Discharge High- 
Pressure Steam Traps For Draining Live-Steam Separators And Reheating-Coils Of 
Two Cross-Compound Engines. 



the trap is intended to discharge. This is accomplished by- 
admitting boiler-steam to the trap. With the equalization 
of pressure, the water which has been collected in the trap 
flows out by gravity. 



406 



STEAM POWER PLANT AUXILIARIES [Div. 13 



489. The Volume, In Cubic Feet, Of Steam Required For 
Each Discharge Of A Return Trap is approximately equal to 
the volume, in cubic feet, of the water discharged. 



Counterweight for Hotting 
Bowl In Filling Position 



Live 
Steam 
Pipe-. 



Hollow Horn of Yoke 

Conveys Steam 
to Live Steam Pipe 



Pipe Conveying- 
Condensation- 
Water to Trap 




Check-Valve, 

Opens Toward 

Trap 



Check -Valve, 

Opens from 

Trap 



Boiler 
Shell 



Pipe Conveying 

< Water from Trap 

to Water Space 
in Boiler 



Boiler 
Check 
-Valve 



Fig. 399.— Bundy Return Trap. When B Fills With Water And Falls, V Opens And 
A Closes. Steam Then Passes Into Bowl Through H And L, And Water Is Forced Out 
Through F, T, C, And D. When B Empties And Rises, V Closes And A Opens. Con- 
densation-Water Then Passes Into Bowl Through W, S, T, And F. 



Example. — Assume that a return trap is discharging into a boiler 
under 100 lb. pressure. Then the weight of the steam, which is admitted 
to the trap is (as taken from a steam table) about 0.25 lb. per cu. ft. 
The returned water of condensation weighs about 60 lb. per cu. ft. Now 



Sec. 490] 



STEAM TRAPS 



407 



as stated, 
of steam, 
steam. 



above, 1 cu. ft. (60 lb.) of water requires 1 
Hence, 1 lb. of water requires 0.25 -j- 60 



Glass 
Wafer 
Gage-- 



fAir Vent .-Copper-Ball Float 
'.- -Toggle -Joint Valve- 
Operating Mechanism 

Plugged Orifice for Ac- 
cess to Seat Bushing-. 
Blowout Vglve- 




cu. ft. (0.25 lb.) 
= 0.0042 lb. of 



., Valve-Seat \ 
Plugged Drainage Orifice-' Bushing--' ouflet- 

Fig. 400. — American Ball-Float Continuous-discharge High-pressure Steam Trap. 



Note. — A portion of the heat of the steam is lost by radiation from the 
trap. Also, steam may be lost, at each discharge, through the vent- 
valve {A, Fig. 399). The cumulative loss from these sources may amount 
to 1 per cent, of the total evaporation of the boiler. 



-Float Rod 



■-Link ..stationary 
.Jaws A pivots 




■■Monel Metal 
Valve Stem fusing. 



Phosphor-Bronze: 
Valve-Seat Bushing-' 



Fig. 401. — Toggle-Joint Valve-Operating Mechanism Of American Ball-Float High- 
Pressure Steam Trap. 



490. The Economy Of Return Steam-Trap Service resides, 
mainly, in the saving effected by returning the water of con- 
densation from high-pressure steam apparatus directly to 
the boilers, instead of returning it thereto in relays, as through 
a receiver or feed-water heater under atmospheric pressure. 

Explanation. — In industrial processes which require steam for heat- 
ing, drying or boiling, the steam is commonly supplied from the boilers, 
and is condensed in the manufacturing apparatus under pressures rang- 
ing from a few pounds up to 100 pounds or more. 

Where steam of, say, 80 lb. pressure is used in heating-coils, as in a 
high-temperature dry-room, the water of condensation may leave the 



408 STEAM POWER PLANT AUXILIARIES [Div. 13 

coils at a temperature of 300 deg. fahr. If such water is trapped to an 
open receiver or feed-water heater, it will, immediately it is discharged 
by the trap, expand and cool to the boiling point under atmospheric 
pressure. Also, its temperature must be still further reduced to about 
210 deg. fahr. in order that its delivery to the boilers, by a feed-pump, 
may be facilitated. Thus the water will have thrown off the heat cor- 
responding to a temperature reduction of about: 300 — 210 = 90 deg. 
fahr. Furthermore, it will have lost a considerable portion of its own 
bulk and some heat through vaporization. While most of the water 
thus vaporized may be recovered, some of it will be a dead loss. 

The saving that might be realized, in this case, by returning the water 
of condensation directly from the heating-coils to the boilers is, there- 
fore, represented by (1) The quantity of coal required to supply the heat 
corresponding io a temperature reduction of 90 deg. fahr. plus (2) The heat 
lost through vaporization. 

491. A Proper Location For A Return Steam Trap (Fig. 399) 
is at least 3 ft. above the normal water-level in the boiler to 
which the trap is attached. This will insure a positive 
gravitational flow of the returned water from the trap to the 
boiler. 

492. The Economy Of Non-Return Steam -Trap Service 
subsists, mainly in the saving effected by preventing the steam 
from blowing through drips and drains directly to the atmos- 
phere. It is contingent upon two principal considerations. 
(1) Selection of the proper type of trap for the particular service 
requirements. (2) The area of the trap discharge-valve orifice 
and the condition of the valve. 

Example. — Where a %-in. drain pipe from a steam-piping system 
under, say, 165 lb. per sq. in. gage pressure is blowing directly to the 
atmosphere, the resulting loss of steam may amount to about 1,120 
pounds per hour. This is the equivalent of, approximately, 35 boiler 
horse power. Assuming that a boiler horse power costs, say, $3.25 per 
mo., the total monthly loss from this source will amount to about 3.25 X 
35 = $113.75. With the drain-pipe connected to a properly-selected 
steam trap, the loss of steam, due to condensation in the drainage 
connections, might be reduced to about 32 pounds per hour. 

Note. — The Area Of The Valve-Orifice Of A Trap For Low- 
Pressure Service should equal the cross-sectional area of the size of 
pipe for which the outlet orifice of the trap is tapped. 

The Area Of The Valve-Orifice Of A Trap For Medium Or 
Ordinary High-Pressure Service, as where the drainage from a live- 
steam separator is discharged into an open feed-water heater, may be 



Sec. 493] STEAM TRAPS 409 

smaller than the openings in the pipe-connections. It should, however, 
in any case, be large enough to obviate liability of the passage becoming 
clogged with particles of scale. 

493. Steam Traps Which Are Dependent Upon Temperature 
Changes For Their Operation Should Not Serve Separators 
Or Similar Apparatus, in the draining of which the trap should 
operate instantly after the accumulated water has attained 
the head at which it should discharge. 

Explanation. — Assume that either a float-operated trap or a tilting 
trap is installed for draining the steam-separator in a supply line which 
ordinarily conveys 90-lb. -pressure steam. The trap will continue to 
function, without intermission, if the pressure rises to, say, 100 lb. or 
falls below 90 lb. But if an expansion trap is substituted, it must, 
necessarily, be set to open at the temperature of the condensation 
from the 90 lb. -pressure steam, which may be as low as 310 deg. fahr. 
Consequently, if the pressure rises to 100 lb., at which the condensation 
may reach the trap at about 320 deg. fahr., the expansion trap will 
remain closed until the temperature of the condensed water drops to 
310 deg. fahr. During the requisite time-interval, however, the con- 
densate accumulation might become dangerously excessive. On the 
other hand, if the pressure falls below 90 lb., condensation may reach the 
trap at some temperature below 310 deg. fahr. Hence, the expansion 
trap will blow steam so long as the diminished pressure continues. 

494. Steam-Traps For Attachment To Heating Coils 

(Fig. 395) and similar apparatus, may, with advantage, 
operate on the principle of thermal expansion. 

Explanation. — Assume that steam at 90 lb. pressure is circulated in a 
set of heating coils. Then the water of condensation will form at a 
temperature of about 330 deg. fahr. Hence, a float-operated trap or 
a tilting trap will discharge the water at approximately this temperature. 
Assuming that the surrounding air is heated to 150 deg. fahr., the 
quantity of heat in the trap-discharged water which will be rendered 
unavailable for radiation from the coils will correspond to a temperature 
range of 330 — 150 = 180 deg. fahr. But if an expansion trap is sub- 
stituted, it may be set to discharge at 150 deg. fahr. Thereby the maxi- 
mum available thermal value of the steam delivered to the coils will be 
realized for heating. 

495. The Proper Location For An Ordinary High- Or Low- 
Pressure Steam Trap (Figs. 394 to 398) is, with reference to 
the location of the apparatus which the trap is intended to 
serve, such that the drainage-water will flow to it by gravity. 



410 



STEAM POWER PLANT AUXILIARIES 



[Div. 13 



Note. — If the apparatus to be drained is located at an inconveniently- 
low elevation, as on the bottom of a narrow pit or trench, an expansion 
trap may be located (Fig. 402) at a higher elevation if the drainage water 
leaves the apparatus under sufficient pressure. There should be at least 
}4 lb. per sq. in. pressure for each foot vertical height. 

Example. — A steam-pressure of 5 lb. per sq. in. in the heating-coil 
(Fig. 402) will, practically, balance a column of water: 5 -5- 0.5 = 10 ft 
high. Hence, the water of condensation will be forced to the trap, if 
the trap-inlet is located less than about 10 ft. above the drainage-outlet 
of the coil. 

Trap Discharge Pipe- 
By-Pass- 




Fia. 402. 



■Method Of Trapping Condensation From Heating Coil Located On Bottom 
Of Deep Pit. 



496. The Location Of An Expansion Trap should be such 
that its operation will not be affected by excessive variations 
of temperature occurring in the surrounding atmosphere. 

497. The Capacity Of A Steam-Trap may be rated (Table 
498) either in terms of the quantity of water to be trapped per 
hour, or in terms of the extent of radiating surface in the ap- 
paratus from which the trap may drain water of condensation. 



Note. — It is commonly assumed that each square foot of direct ra- 
diating surface in a heating system will, ordinarily, condense about 0.33 
lb. of steam per hour. It is also assumed that the radiation from each 
lineal foot of 1-inch pipe in a heating coil will, ordinarily, condense about 
0.19 lb. of steam per hour. Where very wet products are to be dried 
in a kiln or dry-room, a trap for draining the heating coils should be 
selected on a basis of 0.56 lb. of steam condensed per hour per lineal foot 
of 1-inch pipe. Where the heated air is circulated under pressure of a 
fan-blower, the basis of selection should be 0.94 lb. of condensation per 
hour per lineal foot of 1-inch pipe. 



Sec. 498] 



STEAM TRAPS 



411 



498. Table Showing Dimensions And Capacities Of Steam- 
Traps Working Under Medium Pressure (Adapted from 
Swendeman's, A Steam-Trap Catechism). 





Size, in 


Steam 


Diam., in 


in., of 


pressures, 


n., of valve 


pipe 


in lb. 


orifice 


connec- 


per sq. in. 




tions 


(Gage) 



Rated capacities per hour 



Gal. of 
water 
dis- 
charged 



Pounds of 
water 
dis- 
charged 



Lineal feet 

of 1-in. 

pipe 

drained 



Sq. ft. of 
radiating 
surface 
drained 







50 


375 


3114 


5538 


1846 


>* 


H 


75 


459 


3811 


6776 


2258 






100 


530 


4402 


7827 


2609 






125 


593 


4976 


8847 


2949 






50 


584 


4847 


8618 


2873 


•He 


H 


75 


715 


5936 


10554 


3518 






100 


826 


6853 


12184 


4062 






125 


923 


7662 


13624 


4542 




50 


709 


5883 


10460 


3486 


% 


H 75 


868 


7205 


12810 


4270 




100 


1002 


8320 


14793 


4931 






125 


1122 


9302 


16540 


5514 






50 


844 


6998 


12442 


4147 


$& 


1 


75 


1034 


8579 


15754 


5085 






100 


1194 


9986 


17692 


5898 




125 


1334 


11075 


19692 


6564 






50 


1149 


9535 


16954 


5651 


He 


1>4 


75 


1407 


11680 


20767 


6922 




100 


1625 


13486 


23978 


7993 




125 


1816 


15073 


26799 


8933 






50 


1501 


12537 


22290 


7430 


Vi 


1H 


75 


1838 


15252 


27118 


9039 






100 


2122 


17616 


31322 


10441 






125 


2363 


19694 


35017 


11672 



499. The Quantity Of Condensation-Water To Be Trapped 
From A Piping System may be approximately computed by 
the following formula : 

(106) W„ = A f K (lb. per hr.) 

Wherein: W w = weight of condensation in pounds per hour. 
A j = area of piping surface, in square feet. K = conden- 
sation, in pounds per hour per square foot of pipe surface, 



412 



STEAM POWER PLANT AUXILIARIES [Div. 13 



corresponding to the observed steam pressure, as given in 
Table 500. 

500. Table Showing Rate Of Condensation, In Uncovered 
Pipe Lines, Of Steam At Various Pressures. Adapted from 
Elliott Companys' Bulletin G on Steam-Traps. 



Steam pressure, in 
lb. per sq. in. (gage) 


5 

0.7 


10 
0.8 


20 


30 


40 
1.1 


50 


60 


80 


100 


125 


Condensation, in lb. 
per hr., per sq. ft. of 
pipe-surface 


0.9 


1.0 


1.2 


1.3 


1.6 


1.7 


1.9 



Example. — It is found, by computation that the high-pressure piping 
in a boiler and engine plant exposes 2,683 sq. ft. of radiation-area. The 
steam pressure is 115 lb. per sq. in., gage. What size of trap, as listed 
in Table 498 should be used for draining the system? 

Solution. — By Table 500, the condensation rate for steam at 100 lb. 
pressure == 1.7 lb. per hr. per sq. ft. of exposed surface, and for steam 
at 125 lb. pressure = 1.9 lb. per hr. per sq. ft. of exposed surface. Hence, 
the condensation rate for steam at 115 lb. pressure = (1.9 — 1.7) -f- 
(125 - 100) X (115 - 100) + 1.7 = 1.82 lb. per hr. per sq. ft. of ex- 
posed surface. Applying For. (106) W w = A f K = 2683 X 1.82 = 
4,883.06 lb. per hr. Hence, by Table 498 a M~in. trap having a K-in. 
valve orifice should be used. 

501. The Piping Of A Steam-Trap should be adapted to the 
particular service for which the trap is installed. Numerous 
right-angled turns, and runs of excessive length in the dis- 
charge piping, should be avoided. To obviate interference, 
the discharges from low-pressure and high-pressure traps 
should be piped independently. 



Note. — Every Steam Trap Should Have An External By-Pass 
(B, Fig. 394). Also, stop valves, Vi and V 2 , should be inserted between 
the by-pass connections and the inlet and outlet orifices of the trap. 

Strainers In Trap-Inlet Connections (B, Fig. 395) may be used to 
prevent particles of scale, or other solid substance, from entering the 
trap and fouling the valve. 

Provision For Draining Trap Discharge-Pipes, while the traps 
are inoperative, (D, Fig. 398) should be made when the traps are ex- 
posed to freezing in cold weather. 



Sec. 502j STEAM TRAPS 413 

502. Check-Valves Should Be Inserted In The Discharge 
Pipes Of Steam Traps where two or more high-pressure traps 
discharge (Fig. 398) into a common discharge-line or where 
a return-trap (Fig. 399) is used for boiler-feeding. 

Note. — For ordinary high-pressure service, the check-valves (C, Fig. 
398) in the discharge pipes of steam-traps may be of standard weight and 
may be filled with renewable composition discs. For boiler-feed ser- 
vice however, the check-valves (S and Q, Fig. 399) should be extra 
heavy and should have solid brass discs. Check-valves with composi- 
tion discs are ill-adapted to withstand the stresses of boiler-feed service. 

503. A Vent-Pipe Connecting A High-Pressure Trap With 
The Apparatus Drained (P, Fig. 1) is often necessary to 
insure regular operation of the trap. 

Explanation. — With a scant flow of water from the separator (S, 
Fig. 394) the upper part of the trap will contain steam of the same 
pressure as that in the separator. Should a slug of water enter the sepa- 
rator, direct communication between the steam-occupied space in the 
trap and the steam space in the separator will, in the absence of a vent 
pipe, be closed. The flow from the separator will, therefore, cease until 
the steam in the trap condenses. Restoration of an unimpeded flow may 
be further delayed by air mingled with the trapped steam. 

504. The Care of Steam Traps involves periodic inspections 
and, when necessary, repair or replacement of the valves or 
seats. Inspection should be made frequently because the 
flow of water through steam traps cuts into the valves and 
seats, which may then leak or "blow" steam. Since the 
traps are enclosed — as are usually the discharge pipes — a leak 
would not ordinarially be noticed. But by placing the ear to a 
trap, the blowing can, frequently, be detected. A still-better 
method for detecting the leaks consists of providing an opening 
in the discharge pipe, from which the leak is then visible. 
Since, as stated in Sec. 492, losses from leaks readily become 
excessive and expensive, a leaky trap should, immediately, 
be taken from service and repaired upon discovery of the leak. 

QUESTIONS ON DIVISION 13 

1. What are the general uses of steam traps? 

2. What is the distinction between a return trap and a non-return trap? 

3. What types of traps operate on the principle of buoyancy? 

4. Through what media is the expansion principle utilized in the operation of steam- 
traps? (See Fig. 397). 



414 STEAM POWER PLANT AUXILIARIES [Div. 13 

5. What is the essential operating principle of return traps? 

6. What approximate volumetric ratio exists between the water discharged by a 
return trap and the steam required to operate the trap. 

7. What are the apparent sources of loss of heat energy in the operation of return 
traps? 

8. How is the economy of return-trap service principally manifested? 

9. What is the minimum effective elevation of a return trap with reference to the 
boiler it is intended to feed? 

10. What considerations mainly affect the economy of non-return trap service? 

11. Why are expansion traps inadaptable for draining live-steam separators? 

12. What types of traps should be connected to live-steam separators? 

13. Why are expansion-traps well adapted for draining high pressure heating 
apparatus? 

14. What is the proper location for a non-return steam trap relative to the elevation 
of the apparatus it is intended to drain? 

15. Under what conditions might an expansion steam-trap be located above the 
apparatus it is intended to drain? 

16. What are the common bases of rating for steam-traps? 

17. Mention five structural features of general importance in the piping of steam 
traps. 

18. Under what circumstances are check-valves needed in the discharge pipes of 
steam-traps? 

19. Explain the purpose of a vent pipe connecting the top of a steam trap with the 
top of a steam separator. 

PROBLEMS ON DIVISION 13 

1. If a steam trap is 13 ft. above the apparatus to be drained, what pressure will be 
required to force the water up to the trap? 

2. It is found that a certain uncovered pipe line has a surface of 4530 sq. ft. The 
steam in the line is at 80 lb. per sq. in. gage pressure. What size trap should be used? 



SOLUTIONS TO PROBLEMS ON DIVISION 1 
PUMP CALCULATIONS 

1. By Sec. 1, 9 X 22 + 14.7 = 13.5 ft. 

2. By For. (1),P = ^ = ^^ = 13 lb. per sq. in. 

3. Total length of straight pipe = 115 +38 = 153 ft. Three 90 deg. 
elbows = 3 X 8 = 24 ft. of pipe. Two plugged tees = 2 X 16 = 32 ft. 
of pipe. Two globe valves = 2 X 8 = 16 ft. of pipe. Total equivalent 
pipe length = 153+24+32 + 16 =225 ft. Total friction head, 
L hfT = (225 -r- 100) 3.70 = 8.32 ft. head. Head equivalent to 1501b. per 
sq. in., Lhmp = 150 X 2.31 = 346 ft. head. Measured head due to lift, 
Lhmd = 38 ft. head. Total measured head, L hm T = 346 + 38 = 384 
ft. head. Total head on pump, Lht = Lh/T + LhmT = 8.32 + 384 = 
392.32 ft. head (neglecting velocity head). By For. (1) P = L hT -5- 
2.31 = 392.32 -^ 2.31 = 170 lb. per sq. in. 

4. Length of straight pipe = 153 ft. Three 90 deg. elbows =3X6 = 
IS ft. of pipe. Two plugged tees = 2 X 12 = 24 ft. of pipe. Two globe 
valves = 2 X 6 = 12 ft. of pipe. Total equivalent length of pipe = 
153 + 18 + 24 + 12 = 207 ft. = equivalent length of pipe. Total 
measured head = LhmT = 384 ft. Head delivered by pump (Prob. 3), 
LhT = 392 ft. head. Head available as friction head L h /T = L h T — 
LhmT = 392 - 384 = 8 ft. head. Friction head available per 100 ft. of 
pipe =8 4- (207 -=- 100) = 3.86 ft. head. From Table 14, the water 
delivered = about 6}i gal. per min. (This is found by interpolation.) 

5. 90 cu. ft. per m in. = 90 X 7.48 = 673.2 gal. per min. By For. (7), 
di = 4:.95\/V g m/Vm = 4.95V6 73T2/210 = 8.9 in. or a 9-in. suction pipe 
would be selected and 4.95 V673. 2/390 = 6.5 in., or a 7-in. discharge 
pipe would be selected. 

6. By For. (14), V c/ = LAN,/1,728 = 20 X 10 2 X 0.7854 X 65 X 2 - 
1.728 = 118.2 cu. ft. per min. 

7. By For. (17), X = [100(F c/ - V a )}/V cf = [100 X (510 - 487] -v- 
510 = 4.51 per cent. 

8. By For. (18), E, = 100Va/V c/ = 100 X 487 -=- 510 = 95.5 per 
cent. 

, e y Fo , ( 19 ), y. = ^ - » X " 10 ^^" - 

3.9 cu. ft. per min. 

!0. By Fo, (2 0), * , VH? - V 183 ' 35 ^^ 60 = 

3.97, or practically 4 in. 
415 



416 SOLUTIONS TO PROBLEMS 

11. By For. (22): v m = d/L t /di 2 = 5 2 X 80 -h 2 2 = 500 ft. per min. 

12. By For. (23), W u = WL hmT = 20,106 X 38.5 = 774,081 ft.-lb. 

13. By For. (24), P, to = ™± - ^ = 23.5 ** 

14. By For, (30), ft* - ^ - |™*™ - ,0. ,. P . 

i K r, tt ^ -n lOOW^r 100 X 9,000 ,000 X 120 

15. By For. (31), D c = w ~- = ^ = 

30,857,143 ft.-lb. per 100 lb. of coal. 



SOLUTIONS TO PROBLEMS ON DIVISION 2 

DIRECT-ACTING STEAM PUMPS 

1. The effective plunger area is (12 2 X 0.7854) - [(3 2 X 0.7854) -^ 

2] = 109.6 sq. in. The area of opening of each valve is 0.25 X 4 X 

3.14 = 3.14 sq. in. By Sec. (60), 109.6 X 0.3 -r- 3.14 = 10.5, or, prac- 
tically, 11 valves. 

SOLUTION TO PROBLEMS ON DIVISION 3 
CRANK-ACTION PUMPS 



1. Substituting directly in For. (48) : 
V gm L hmT 150 X 225 



26 h. p. (Use 25 h. p. motor) 



1,300 1,300 

2. V gm (For. 49) = 30 X 0.9 = 27 gal. per min. 
L hmt = 50 + 175 ft. = 225 ft. 
K (Table 108) =0.69 

L f K = 0.69 X 175 = 121 ft. 

Substituting in For. (49), P bhp = ^^coV^ = 4 - 7 h - P- 

(Use 5-h.p. motor). 

SOLUTIONS TO PROBLEMS ON DIVISION 4 

CENTRIFUGAL AND ROTARY PUMPS 

1. By For. (51), the velocity = v m = 481 y/hj = 481 X V160 = 
6,085 ft. per min. 

2. The circumference of the impeller = 6,085 -J- 1710 = 3.558 ft. or 
3.558 X 12 = 42.7 in. The diameter = circumference -r- 3.1416 = 42.7 
-=- 3.1416 = 13.6 m. 

3. By For. (53), the head produced at the new speed = Luti — 



^■=(® 2 x 90 = 177 * 



SOLUTIONS TO PROBLEMS 417 



4. By. For. (52), the quantity of water delivered at the new velocity = 

2,520 X Pwp 



N 2 XV„ m i 1,600 X 400 
ffm2 = ^— - = YJ50 = 441 ( J aL P er mm 



5. By For. (61), the ividth of a single belt = L w 
4 in. 



N Xd 
2,520 X 10 



900 X 7 

SOLUTIONS TO PROBLEMS ON DIVISION 5 

INJECTORS 

From For. (62) 

. _ xH v + (Tjs - T fd ) 

i. w Stt - ^ _ Ti 

Here T/* = 60 deg. fahr. and T fd = 200 deg. fahr. For 100 lb. 
per sq. in. gage the following valves are found in the steam tables: 
Tf S = 338 deg. fahr., H v = 879.9 B.t.u. per lb. When there is a moisture 
content of 2)4 per cent., x = 1.00 - 0.025 = 0.975. Then substituting 
in For. (62): 

0.975 X 879.9 + (338 - 200) „ , , „ . , „ 

W SVJ = ^r^ x^j =7.11 lb. of water pumped per lb. 

of steam 
Again applying For. (62) : — 

. in _ 0.975 X 879.9 + (338 - T fd ) 
2 * 1U T fd - 60 ~~~ 

Transposing and simplifying: 
10T/ d -600 = 857.9+338-!r /d 
11 r /d = 1,796 

Therefore : 

T fd = 163.26 deg. fahr. 

3. From For. (69), for water tube boiler: 

Gallons per hour of injector = — ^l = — — = 206.7 

Increasing by 30 per cent, these results: 206.7 + (30 X 206.7) = 
268.71 gal. per hr. 

Looking in Table 194, the Size B is required to pump 260 gal. per hr. 

Therefore it is the size to use. Note. If the lift is very great (over 15 
feet) it is advisable to select the next larger size of injector. 

4. From Table 194, under Pipe Connections the size given is Y± in. for 
the injector Size B of Prob. 3. This is the correct size for all steam and 
delivery lines, except when the run is unusually long. The suction line 
will be % in. for an 8 ft. lift. For a 15 foot lift, a 1-in. suction line would 
be recommended. For a lift of 20 ft., it would be advisable to use 134-in. 
pipe for the suction line. 

27 



418 SOLUTIONS TO PROBLEMS 

SOLUTION TO PROBLEMS ON DIVISION 6 
BOILER FEEDING APPARATUS 

1. By For. (74) gal. per hr. required = 6 X V Bhp = 6 X 600 = 3,600 
gal. per hr. 

To retain the same per cent, excess capacity when boilers are forced 
225 per cent, the capacity is: 

3,600 X 2.25 = 8,100 gal. per hr. 

2. Pounds of water per hour required by main engine = 

500 X 33,000 X 60 v 1 _ _ . _ Ann „ . 
150>000>000 X 1,000 = 6,600 lb. per hr. 

The auxiliaries require 10 per cent, of this or 660. 

The total normal requirement is then 6,600 + 660 = 7,260 lb. per hr. 
A 50 per cent, excess over this capacity = 1.5 X 7,260 = 10,890 lb. per 
hr. 

There are about 8.34 lbs. of water in a gallon. 

Therefore the pump capacity in gallons = o oa ~ 1>305 gal. perhr. 

SOLUTIONS OF PROBLEMS ON DIVISION 7 
FEED- WATER HEATERS 

1. The quantity of exhaust steam used in heating the feed- water = 
(500 X 20 X 11) 4- 100 = 1,200 lb. per hr. The total heat in the steam, 
above 32 deg. fahr., is about 1,150 B.t.u. per lb. Hence, by For. (77): 

_ r /1 W / + 0.9W.(g +32) 
1/2 ~ W/ + 0.9 W s 
(90 X 10,000) + [0.9 X 1,200 X (1,150 + 32)] 1ft _ . , , . 
1Q?00 o + ( o. 9 x 1?200 ) = ™^ deg. fahr. 

2. As given in a table of the properties of saturated steam, the total 
heat, above 32 deg. fahr., in steam at 150 lb. persq. in., gage, is 1,195 B.t.u. 

rp rp 

per lb. Hence, by For. (76), the saving = H f = y? — ! -jw> ^W; 100 

"■ ~ \ 1 n ~ 3z; 

° 1,195- (W - 32) X 100 - 12 " 85 ^^- 

3. As given in a table of the properties of saturated steam, the total 
heat, above 32 deg. fahr., in steam at 125 lb. per sq. in., gage, is 1,192 B.t.u. 
per lb. Hence, by For. (76) the probable thermal saving = H f = 

m rp 212 150 

B -'hn = rW) 10 ° " U92 -(150 - 32) X 10 ° = 5 ' 8 ^ cent. The 
present annual cost of the fuel supply = 3.5 X 5 X 300 = $5,250. 
Hence the probable annual saving = - — ^Tu) — "~ = $304.50. The 
interest on the investment = — ^j — = $18. The annual cost of 



SOLUTIONS TO PROBLEMS 419 

depreciation = 1QQ = $18.00. The annual cost of maintenance 

and operation = 12 X 4 = $48. Hence, the probable annual net 
saving = 304.50 - (18 + 18 + 48) = $220.50. 

4. By Table 278, U = 350. Hence, by For. (81) A f = ^ f{Tf2 ~ 1 * r-, 

v v» — 2—; 

15,000 X (200 - 70) 
= 35 0x(220- 7 A^y =65 - 6S? - /i - 

5. By For. (78) the weight of steam condensed, 

{T n - T fl )W f = 

W s = 0.9(H + 32) - T f x + 0.1 T/2 

(205 - 40)15,000 



0.9(1,150.4 + 32) - 40+ (0.1 X 205) 



2,370 lb. per hr. 



SOLUTIONS OF PROBLEMS ON DIVISION 8 
FUEL ECONOMIZERS 

1. By Fig. (266), the weight of combustion gases which contain 
12 per cent, of C0 2 = 13 lb. per lb. of coal. Also, the weight of combus- 
tion gases which contain 8 per cent, of C0 2 = 20 lb. per lb. of coal. 
Hence, the leakage-air = 20 — 13 = 7 lb. per lb. of coal. The heat, 
above the outside air-temperature, which is contained in the gases 
leaving the boiler = (550 - 50) X 13 X 0.24 = 1560 B.t.u. per lb. of 
coal. The heat, above the outside air-temperature, which is contained in 
the leakage air =. (250 - 50) X 7 X 0.24 = 336 B.t.u. per lb. of coal. 
Hence, the per cent, of loss = 336 -s- 1,560 = 21.6 per cent. 

2. By For. (82) the requisite ratio = X = T fg -i- T fw = C w W w /C g W 
= (1X8)t (0.24 X 15) = 2.22. 

3. By Fig. 276, the lowest temperature-difference consistent with 
profitable operation = 100 deg. fahr. Hence (Sec. 305) the lowest 
permissible temperature of the gases leaving the boiler = 100 + 377.5 = 
477.5 deg. fahr. By Fig. 277, the corresponding boiler heating-surface 
= 9 to 10 sq. ft. per h.p. 

4. By Fig. 278, the least temperature-difference, consistent with 
economy, between the water and the gases = 40 deg. fahr. Hence, the 
lowest permissible temperature of the gases at exit = 200 + 40 = 240 
deg. fahr. 

5. The heat-transfer = 5.5 X 300 = 1,650 B.t.u. per sq. ft. per hr. 
Hence, the requisite area of heating-surface = 50,000 X 50 -5- 1,650 = 
1,515 sq.ft. 

6. By For. (83), X = 100 {T" fw - T' fu )/{H + 32)-T' /u , = 100 X 
(250 - 110) -h (1197.3 + 32 - 100) = 12.4 per cent. 

7. The annual cost of fuel without the economizer = (2,400 X 24 X 
4.3 X 300 -^ 2,000) X 4.25 = $157,896. The saving effected by the 



420 SOLUTIONS TO PROBLEMS 

economizer = 157,896 X 12.3 + 100 = $19421.21. The annual cost of 
operation, maintenance and depreciation of the economizer = 12,000 X 
0.15 = $1,800. Hence, the net annual saving = 19421.21 - 1,800 = 
$17,621.21. 

SOLUTIONS TO PROBLEMS ON DIVISION 9 

STEAM CONDENSERS 

1. By For. (84), the greatest possible thermal efficiency non-condensing = 
„ Ti - T 2 (450 + 460) - (255 + 460) 125 _ . _ 

Et = -TT - = 450 + 460 = 910 = 2 ° *" cent ' 

Also by For. (84), the greatest possible thermal efficiency condensing = 

T x - T 2 (450 + 460) - (80 + 460) 370 
* = — T — = + 450 + 460 " - 910 = 4 °- 7 Per CmL 

2. By For. (85) the saving in power due to condensing operation = 

4QP hmv 49 X 26.5 ,_ _ 

— 5 — = == = 16.6 per cent. 

l m to 

3. By the graph, Fig. 286, the ideal steam consumption of the turbine 

is 14 lb. per h.p. hr. non-condensing and 7 lb. per h.p. hr. condensing. 

The actual steam consumption condensing, then = 

7 
-j X 22 = 11 lb. per h.p. hr. 

4. The absolute condenser pressure = 29.8 — 27 = 2.8 in. of mercury. 
By For. (87), the absolute pressure = 

t-j Lhmb — Lhmv 29. 2ii ., 00 ,, 

P ' = 2.03 = ^703- = ° 8 lh - per Sq - m - 

The per cent, of the possible vacuum = 

27 

7 r~ X 100 = 90.6 per cent. 
Zv.o 

5. By For. (88), the volume of the condenser = V = 0.001,43 W g + 
8.25 = (0.001,43 X 10,000) + 8.25 = 22.55 cu. ft. One hour = 3,600 
sec. One cubic foot of water weighs 62.5 lb. Therefore, the volume of 
the cooling water = 

10,000 X 36 



3,600 X 62.5 



= 1.6 cu. ft. per sec. 



The volume of the condensate = q fuvw fi9 r. = 0-044 cu.ft. per sec. 
Hence, the tail-pipe area = ^^^ = 0.329 sq. ft. = 0.329 X 

/47.4 
144 = 47.4 sq. in. Therefore, the tail-pipe diameter = \/-^^ or 7.8 in. 

or approximately, 8 in. 



SOLUTIONS TO PROBLEMS 421 

6. By Table 345, the steam temperature corresponding to a 27 in. 
vacuum = 115.06 deg. fahr. Hence, the temperature difference between 
the discharge and the entering steam = 115.06 — 105 = 10.06 deg. fahr. 

By Table 345, total heat in steam at a 27 in. vacuum = 1110.2 B.t.u. per 

lb. By For. (89), the weight of cooling water required = 

... _. tf-T/o + 32 innnn 1110.2 - 105 + 32 , 1 . OQf . 7 , 

W "= W * Tn L Tfx = 10 ' 000 105 - 80 = 414 ' 88 ° lb - per 

hr. One gallon = 8.3 lb. of water. Hence the volume of cooling-water 

. 414,880 CQQ _ 

required = pn v o o = °33 gal. per mm. 

7. As referred to a 30-in. barometer (Table 345) the degree of vacuum = 
30 - 29.5 + 28 = 28.5 in. of mercury. By Table 345, the total heat of 
steam in. a 28.5 in vacuum = 1,100 B.t.u. per lb. By For. (89), the weight 
of water required, 

W - = ™- H T /2 T -Tn 2 = 10 - 00o l ' 10 8 °7~- 8 67 +32 = 523 - 500 ">■ P er hr - 

8. By For. (90), the quantity of heat to be abstracted from the steam, 
H t = W S (H - T fc + 32) = 150,000(1095.6 - 80 + 32) = 157,000,000 
b.t.u. per hr. By For. (91) the tube surface required, 

f u{T f . - y 2 [T fl + r,d) 

U, by Table 350 = 600. T fs , by Table 345, = 82 deg. fahr. Hence, 
. 157,00 0,000 10 .__ ft 

Ai = 600(82-^[60 + 77]) = 19 ' 4 °° ^ fL 

SOLUTIONS TO PROBLEMS ON DIVISION 10 
METHODS OF RECOOLING CONDENSING WATER 

1. By Table 393, the relative humidities of the air at entrance and 
exit are, respectively, 55 per cent, and 92 per cent. By Fig. 315, the 
weight of saturated water vapor per cubic foot of air = 0.001,1 lb. at 
70 deg. fahr. and 0.002,2 lb. at 90 deg. fahr. Therefore, the moisture 

content of the air at entrance = — —=: = 0.000,605 lb. per cu. ft., 

0.002 2 X 92 
and of the air at exit = ~ ~: = 0.002,024 lb. per cu. ft. Hence, 

the quantity of water absorbed per cubic foot of air = 0.002,024 — 0.000,605 
= 0.001,419 lb. 

2. By Sec. 402, the area required for a simple cooling pond = 1,000 X 
120 = 120,000 sq. ft. 

By Sec. 411, the area required for a spray pond = — ^~ — : = 

3,000 sq. ft. 

3. By solution of Problem 1, the quantity of water evaporated, per cubic 
foot of air-flow through the tower, = 0.001,419 lb. By Sec. 399, the heat 
abstracted, per pound of water evaporated, = 1,000 B.t.u. 1 gal. = 8.3 lb. 



422 SOLUTIONS TO PROBLEMS 

,„,' , ,, ., f . n . t 800 X 8.3 X 20 X 80 

Therefore, the volume of air-flow per minute = 00 q 1419x1 ooq x 100 = 

74,870. cu. ft. 

By For. (92), E = 100 ^ ~ %£ = 100 X ^ ~ ^ = 44.4 per cent 
The water lost by evaporation = 74,870 X 0.001,419 = 106.24 lb. per 

min. or onn v o o X 100 = 1.6 per cent. 

4. By solution of Problem 3, the volume of air-flow = 74,870 cu. ft. 

per min. By Sec. 423, the allowable velocity of air-flow = 700 ft. per min. 

74 870 
Therefore, the/ree area = 7 ' = 107 sq. ft. 

By Sec. 423, the free area = 64 per cent, of the total horizontal cross- 
sectional area. Therefore, each side of the base = *v/ ^7 = 

13 ft. 

5. By Table 408, the discharge, per nozzle, = 60 gal. per min. or 60 X 
60 X 24 = 86,400 gal. per day. Hence, the requisite number of nozzles = 
40,000,000 

86,400 ^ 6 - 

By Sec. 411, 1 sq. ft. of pond area will suffice to cool 250 lb. of water per 

hour. 1 gal. = 8.3 lb. Hence, the requisite area = — ' ~ ' ~ 7 — — = 
55,333 sq. ft. 

SOLUTIONS TO PROBLEMS ON DIVISION 11 
STEAM-PIPING OF POWER-PLANTS 

1. A table of the properties of saturated steam shows the density at 
150 lb. pressure, gage, to be 0.363 lb. per cu. ft. By For. (97) di = 

13.54 \/ Q ' ' Q00 = 5.6 in. or practically 6 in. 

2. By For. (99), d im = Vdn* + d i2 2 + drf + etc. = 

V2.5 2 + 4 2 + 5 2 + 7 2 = 9.8-m., or, practically, 10-in. 

3. By For. (100), L v = 114 d { -i- (l + ^£\ = 114 X 6 +- (l + ^f) = 

427.5 in. or 427.5 ^ 12 = 35.6/*. By For. (101) L e = 7Qd { + (l + -£ J 

= 76 X 6 -r- (l + -g-j = 285 in. or 285 4- 12 = 23.75 ft. Hence, the 

total equivalent pipe-length = 35.6 X 2 + 23.75 = 94.95/*. 

4. A table of the properties of saturated steam gives the temperature 
at 135 lb. pressure, gage, as 358.5 deg. fahr. A manufacturer's table of 
pipe sizes (Nat. Tube Co.) gives the outside diam. of an 8-in. pipe as 
8.625-in. By For. (103), L b = 0.04S\/doL p T f = 0.043 X 

V8.625 X 150 X (358.5 - 60) = 26.72/*. 



INDEX 



427 



Page 
Centrifugal pump for condenser cir- 
culating pump, advantages 308 

for condensing water 135 

foundation bolts for 140 

foundation for 140 

guide bearing, illustration. . . . 122 

handling hot water 137 

head, volume, R.P.M. rela- 
tions, graph 126 

heads and speeds 126-128 

illustration of principle 103 

impeller, theoretical speed. . .. 106 
in connection with feed-water 

heater 136 

independent suction lines for, 

illustration 143 

installation 138 

left-hand, illustration, defini- 
tion 140 

leveling 141 

maintenance 151 

mechanical efficiency 193 

methods of driving 131 

methods of priming 147 

motor-driven, for boiler feed 192 

speed variation 151 

speeds 132 

multi-stage, for high heads ... 112 
open-type impeller, illustra- 
tion 118 

operating principle 101 

performance characteristics. . . 122 
positions of discharge-nozzles. 140 
power required and speed of 

impeller, formula 108 

power required to drive at any 

speed 130 

pressure head and speed of 

impeller 108 

priming 146 

put in service, bearings 

cleaned 151 

quantity of water delivered 107 

required belt width 131 

right-hand, illustration, defini- 
tion 139 

run in wrong direction 151 

with discharge valve closed 149 

with casing empty 147 

selection of 138 

single- and double-suction. . . . 114 
single-stage volute, illustra- 

. tion 103 

size and capacity 107 

speeds heads and capacity. 126-128 

starting 149, 151 

submerged type 120 

submersible type, illustration. 150 
suction lift of, how measured . 8 

piping 141 

vacuum pipe 145 

test 122-124 

test conditions for 124 

theory 101 

two-stage double-suction tur- 
bine, illustration 110 

explanation 114 

vertical shaft 120 

submerged, thrust-bearing 

for 152 

thrust bearing for 121 

volute and turbine, applica- 
tions of Ill 

water rate of turbine for 193 

with branch discharge pipe, 

illustration 149 

separator 389 



Page 

Centrifugal separator, operating prin- 
ciple 390 

Chamber, vacuum, in pump, func- 
tion of 51 

Check valve, faulty boiler 169 

in centrifugal pump dis- 
charge 146 

in discharge pipes of steam 

traps 413 

Checkerwork, cooling tower, com- 
puting height 360 

cypress board, illustration 351 

for mixing chamber 352 

wood, for cooling towers 353 

Circulating pump, surface condenser 321 
Cleaning condenser, cost, what 

determines 324 

jet and surface condensers, 

relative cost 323 

Coal consumption, duty of steam 

pump on basis of, formula 32 

with economizer, graphs 271 

"Coal Miner's Pocketbook" on pump 

management 66 

Cochrane feed-water heater, illustra- 
tion 226 

heater in condensing plant, 

illustration 180 

horizontal receiver - separator, 

illustration 387 

open induction feed-water 

heater 219 

Coefficient of heat transference, 
value effected by conditions 

in condenser tubes 304 

Coke heater packing 246 

Column, fluid, static head definition 
of water, converting to unit 

pressure, formula 4 

Combining tube of injector 156 

Combustion-gas, temperature, loss 
of, ratio to gain of feed- 
water temperature 266 

Commonwealth Edison Company, 
Fisk Street Station, boiler 
and economizer surfaces 269 
Condensate from heating, traps for 

returning 203 

issuing from jet condenser, 

velocity 299 

Condensation and entrainage in 

steam-piping, steam trap for 403 
due to loss of heat from bare 

steam pipes formula 381 

from heating coil, method of 

trapping, illustration 410 

in high-pressure steam-piping 
returned to boilers with 

Holly loop 383 

in primitive steam engines 278 

in steam-heating and power 

apparatus, steam trap for . . 403 
rate of steam at various pres- 
sures in uncovered pipe 

lines, table 412 

Condensation-water, quantity to be 
trapped from piping system, 

formula 411 

Condenser, Alberger barometric, dry 
air pump connection, illus- 
tration 309 

ammonia, condensing water for . 329 
low-temperature cooling tower 

test data, table 358 

temperature of water leaving 332 
and spray-cooling outfit, per- 
formance guarantees 347 



428 



INDEX 



Page 

Condenser, application of, F. A. Burg 316 

auxiliary, types of pump used as 305 

Baragwanath single flow 294 

barometric jet, adjustment and 

care 315 

Buckley siphon, diagram 292 

care of.... 312-316 

classification 289 

cleaning 315 

cleaning, what determines cost 324 
cooling water character, quan- 
tity and source factors in 

selection of 318 

counter-current 289 

double-pipe ammonia, with 
"Burhorn" metallic cooling 

tower 330 

dry-tube 300 

ejector jet, adjustment and care 315 

operation 298 

elementary jet, illustration 289 

volumes of air, water and 

steam, diagram 283 

evaporation cooling 290 

feed-water quality factor in 

selection 318 

for steam-driven prime mover, 

selection of 325 

increase in thermal efficiency 

due to, formula 281 

jet 289 

and surface relative cost 323 

connected in parallel 296 

effect of bad water 319 

ejector 289 

first cost less than that of 

surface 323 

how to restore vacuum and 

condensing operation 314 

low-level 289 

or surface, quantity of cooling 

water required, formula .... 299 
power requirements compared 

to surface condenser 322 

pumping head 320 

pumping head of circulating 

pump 321 

pumps, cost of maintaining . . . 323 
ratio of water to steam fixed 

for given vacuum 321 

requisite size, formula 298 

table of operating costs 326 

temperature of water dis- 
charged from 300 

two on same exhaust line 296 

velocity of condensate issuing 

from 299 

velocity of exhaust steam 

entering 299 

with reciprocating engine .... 296 
joints and stuffing boxes, index 

to condition 316 

Koerting multi-jet ejector, illus- 
tration . 293 

leakage of air, prevention of ... . 312 

low-level jet, operation 295 

most profitable vacuum in 285 

multi-flow 293 

illustration 294 

of compound condensing engine, 

cooling tower data table ... 361 

parallel current 289 

power saving due to, formula. . . 281 
power to remove air and water 

from 284 

pump, electric or steam, type of 

drive for 322 



Page 

Condenser pumps 305 

selection and economics. . . . 316-318 
selection for given installation . . 316 

siphon jet 289 

or barometric jet, operation . . 297 
without pump, apparatus for 

starting 298 

jet, started and operated with- 
out pump 298 

standard jet, starting 295 

stopping 297 

steam 277-327 

condensing water for 329 

definition, purpose 277 

high-temperature cooling 

tower test data table 359 

saved by 279 

saving due to, graph 282 

temperature in 330 

surface 289 

air-cooling 290 

and jet, comparison of pump- 
ing heads 320 

built-in-tube-cleaning equip- 
ment. 324 

circulating pump pumping 

head 321 

discharge from used for boiler 

feed-water 388 

double flow 292 

effect of bad water 319 

effectiveness preserved by 

exhaust-steam separation 388 

elementary, illustration 291 

first cost greater than that of 

jet 323 

forced-draft cooling tower 

with 354 

fouling of tubes, result 315 

heat transference in 300 

leaky tube ends, result 315 

LeBlanc 311 

operation 300 

per cent, of boiler feed returned 

to boiler 319 

power requirements compared 

to jet condenser 322 

pumping head 320 

pumps, cost of maintaining. . . 323 
purity of water delivered to 

feed heater 318 

ratio of water to steam varied 

to suit conditions 321 

replacement of tubes 325 

single flow 292 

table of operating costs 326 

temperature "drop" in, defini- 
tion . 304 

typical, illustration 291 

water-cooling 290 

Westinghouse-LeBlanc 311 

temperature, vacuum corre- 
sponding to 315 

tubes, metals used for 302 

tube gland, Worthington stand- 
ard 302 

grease removed from 3l6 

heat transfer in . , 304 

sizes 302 

vacuum, gages for measuring . . . 284 

loss while running 313 

relation to steam consumption 286 

Watt, engine with 278 

Wheeler rain type low-level jet 294 

when hot will not work 314 

Worthington independent jet, 

illustration 290 



INDEX 



429 



Page 
Condensing equipment, arrange- 
ments 341 

operation, advantages 288 

table snowing economy 287 

work gained by, diagram 280 

plant, steam useful for feed- 
water heating 230 

plants, vacuum and atmospheric 

feed- water heaters 217 

water, centrifugal pump for .... 135 
for steam and ammonia con- 
densers 329 

recooling methods 329-361 

Condensing-engine exhaust line, 
vacuum trap for draining 

separator 404 

Conduction, cooling by 329 

Connecting rods for deep-well pump 88 

Contra-flow in economizer 265 

Convection, cooling by 329 

Cooling, atmospheric, limit of 333 

device, atmospheric, efficiency, 

formula 334 

pond area data 342 

cooling effect on condensing 

water 329 

depth 343 

diagram 337 

evaporation rate formula 337 

requisite surface area 340 

satisfactory for small plants . . 337 
simple, requisite area com- 

' puted. •••..-... 342 

spray fountain with 343 

tower, atmospheric, average 
temperature reduction in 

summer, formula 356 

classification 353 

closed, loss of water less than 

in open 355 

closed or chimney-flue, forced 

draft 354 

closed or chimney-flue, natural 

draft 353 

closed or flue completely 

enclosed 355 

closed, using either forced or 

natural draft 354 

" Cooling Tower Company's Cata- 
logue," formulafor tempera- 
ture reduction in summer . . 356 
temperature reduction 

formula 347 

design, distributor and 

decks, illustration 351 

high temperature natural 
draft cooling tower test 

data, table 359 

"impact " nozzles 343 

impact spray nozzle 345 

installation, spray-nozzle 

tests, graph 346 

low temperature cooling 

tower test data, table .... 358 
on atmospheric cooling .... 333 
table of atmospheric condi- 
tions 331 

computations based on tests 

and practice 355 

principles involved 355 

construction and operation . . . 350 
cooling effect on condensing 

water 329 

cost 361 

fan-blower height required . . . 361 

for artificial cooling 318 

forced-draft, for condensers 



Page 
of compound condens- 
ing engines, data table. 361 
with surface condenser. . . 354 
high-temperature natural 

draft, for steam condenser, 

test data, table 359 

low-temperature natural draft, 
for ammonia condensers, 

test data table 358 

open, loss of water greater 

than in closed 355 

open or atmospheric 353 

per cent, of recooling resulting 

from evaporation 352 

performance, typical data. . . . 357 
proportions, method of com- 
puting 360 

total height 361 

water-loss per cent 357 

wood checker work for 353 

Worthington, illustration 339 

water, character, quantity and 
source, factors in condenser 

selection 318 

cost of handling 320 

required for jet condenser. . . . 299 

Corrosion of economizers 257 

Corrosive liquids, pumping 93 

Cost of operation of boiler-feed pump 179 

Counterflow principle 300 

Coupling, flexible, for connecting 

centrifugal pump 133 

illustration 134 

Crane Company, working pressures 
for wrought iron and steel 

pipes, data 367 

Crank-action power pump 75 

classification 82 

compared with direct-action. . 78 

data, table 93 

for deep well, classification ... 85 

piston speeds, table 92 

walking-beam type 76 

water-ends of 80 

Crank-and-fly-wheel pump 75 

advantages and disadvantages 80 

application 95 

economies 77 

hydraulic elevator, illustra- 
tion 78 

in city water works 79 

sizes 94 

steam pumps 79 

Cross-head of duplex-pump 62 

Cross-heads secured to duplex-pump 

piston-rods 61 

Cup-washers for deep-well pump- 
plungers 98 

for pump-plungers, table of 

dimensions 99 

Current-flow, relative, through econo- 
mizer 265 

Cushion-valves for duplex-pumps ... 64 

Cylinder area, definition 34 

area of water, formula for 35 

diameter of water pump, formula 

for 35 

steam, area to balance given 

water pressure, formula .... 37 
water, area to balance given 

steam pressure, formula 37 

Cylinder-head, marking striking 

point 62 

D 
Dead-center, avoided by valve-stem 

lost motion 58 



430 



INDEX 



Page 
Decks and distributor of Cooling 
Tower Company design, 

illustration 351 

Deep-well pump, details of 87-89 

pumps 84 

DeLaval Steam Turbine Company, 

piston balancing system. ... 117 

Delivery pipe, friction-head in 14 

tube of injector 156 

Delivery-lift of the water 7 

Diagram, indicator, of direct-acting 

steam-pump 25 

Diaphragm for pump governor 202 

Differential piston for deep-well 

pump 94 

Diffusion vanes 109 

Direct Separator Company on sepa- 
rator economy 387 

Discharge, actual, of pump 21 

level of condenser higher than 

circulating pump 320 

line from injector 167 

of piston or plunger pump, 

formula 22 

pipe of pump, sizes for, formula 18 
rate of, crank-action pump, 

graph 91 

theoretical, of pump 21 

Discs, seats of, pump-valves 45 

Displacement of plunger pump, 

formula 19 

of pump units of 19 

of reciprocating pump of 19 

Distribution of steam from boiler 

plant, methods for 371 

Distributor and decks of Cooling 
Tower Company design, 

illustration 351 

Double-suction pump 114 

Draft, artificial, with economizer. . . . 263 

available for economizer 273 

forced in cooling tower, advan- 
tages and disadvantages . . . 355 
in chimney in inches of water, 

table 263 

natural, economizer with 262 

natural, in cooling tower, ad- 
vantages and disadvantages 355 
Draft-pressure drop through econo- 
mizer 262 

Drainage, pumps for 135 

Draining, automatic, live-steam 

separator 400 

Drive, type for condenser pumps .... 322 
Driving horse-power, total, steam- 
pump, definition 29 

Driving unit selected for pump 138 

Drop, temperature, in surface con- 
densers, definition 304 

Dry-tube condenser 300 

Dry-vacuum pump, illustration 79 

Duplex boiler-feeder 204 

double-acting power pumps, 

application 96 

fire-pump, illustration , , 27 

single-acting pumps, application 95 

Duplex-pump, air-chamber for 50 

compared with simplex-pumps . . 64 

compound, illustration 65 

compression-space 63 

cross-heads secured to piston- 
rods 61 

for high-pressure service 65 

outside adjustment of lost 

motion .• • • ■ ^3 

pistons, steam-cushioned, illus- 
tration 63 



Page 
Duplex-pump, slow-running, valve- 
stem lost-motion 63 

Duty of pump, basis of steam con- 
sumption, formula 33 

of pump on coal basis definition 32 
of steam pump, coal consump- 
tion, formula 32 

of steam pump on basis of heat 

consumed, formula 33 

Dynamic head or pressure, definition 5 

E 

Eclipse exhaust-steam separator, 

illustration 400 

]r; c onomizer, advantages 272 

by-pass, installation, illustra- 
tion 261 

cast-iron, wrought iron and 
steel, advantages and dis- 
advantages 257 

coal consumption with, graphs. . 271 
compared to additional boiler 

heating-surface 267 

contra-flow 265 

characteristics, graph 266 

installed, example 269 

corrosion of 257 

cost of installation 275 

definition 210, 251 

disadvantages 272 

draft-pressure drop through .... 262 
forced draft and induced draft 

fans 262 

fuel 251-275 

fuel-saving due to, formula ..... 270 

heat transfer in 269 

heat utilized in 252 

heating-surface, least tempera- 
ture-difference for 269 

ratio to boiler-horsepower. . . . 269 

high- and low-pressure 254 

independent 252 

construction 254 

illustration 253 

initial cost 273 

inspection 274 

installation, conditions deter- 
mining 273 

integral 252 

integral, high- and low-pressure. 255 

leakage of air into 260 

minimum temperature difference 266 

parallel-flow 265 

characteristics, graph 266 

ratio of loss of combustion-gas 
temperature to gain of feed- 
water temperature 266 

relative current-flow through . . . 265 

scale forming in 259 

setting, air infiltration through 260 

steam-generating efficiency 271 

surfaces, pitting on 274 

table of temperatures obtained 264 

tubes, arrangement 255 

clean, saving due to, graph . . . 259 
in staggered rows, illustration 256 
in straight rows, illustration. . 256 

method of removing soot 257 

scale and sediment in 259 

sweating 257 

tube-surfaces, cleanliness 257 

types of 252 

with artificial draft 263 

with natural draft. 262 

Efficiencies, relative, boiler-feeding 

devices 173 



INDEX 



431 



Page 
Efficiencies, total, steam pumps, table 31 
Efficiency, hydraulic, data necessary 

to determine 28 

hydraulic, of pump, formula .... 27 
indicated, reciprocating pump, 

formula 26 

maximum mechanical direct- 
acting steam-pump 30 

mechanical, reciprocating pump, 

definition, formula 30 

of pump, causes impairing 66 

thermal, increase due to con- 
denser, formula 281 

total, steam-driven pump 31 

of pump, definition, formula. . 30 
values for different pumps. ... 31 
volumetric, of pump, definition, 

formula 22 

volumetric, relation to pump- 
slip 22 

Ejector jet condenser 289 

adjustment and care 315 

or jet pump as condenser aux- 
iliary 308 

priming 148 

for pump, illustration 146 

Elbow in pump piping 6 

"Electric Journal," F. A Burg, Ap- 
plication of Condensers. . . . 316 
motor drive, centrifugal pumps 132 
pumps, advantages and dis- 
advantages, table 189 

Electrically-driven pump, advan- 
tages 94 

Electric-drive for boiler-feed pump . . 182 
Elevator, hydraulic, crank-and-fly- 

wheel pump for 78 

pumps for 135 

" Elliott Companys' Bulletin G" on 
Steam-Traps, condensation 

rates table 412 

Engine, condensing, power devel- 
oped 279 

steam consumption table 287 

cylinder, water in, danger 386 

economy increased by condenser 

creating partial vacuum .... 277 
non-condensing, power devel- 
oped 279 

piping, steam separators in, 

illustration 385 

primitive steam, condensation 

in 278 

reciprocating, condenser vacuum 

for 285 

jet condenser with 296 

steam or gasoline, driving cen- 
trifugal pump 133 

Watt's condensation, illustra- 
tion 279 

double-acting condensing, 

illustration 279 

with Watt condenser 278 

Engines, several, selecting separators 

for 400 

Entrance-head, definition 6 

Equalizer pipes, function 372 

Evaporation cooling in condensers 290 

heat carried away 356 

extracted in 340 

of water, cooling effect on con- 
densing water 329 

rate, cooling pond, formula 337 

weight, formula 339 

Excelsior heater packing 246 

Exhaust heater, effect on boiler 

feeding efficiency 174 



Page 

Exhaust heater, sound muffled by ex- 
haust head 399 

steam available, effect on need 

for economizer 273 

energy in, non-condensing 

plant 207 

entering jet condenser, veloc- 
ity 299 

for feed-water heating 211 

from engine, heat in 230 

heat available in 209 

intermittent delivery to feed- 
water heater 242 

main piping, average pressure- 
drop 376 

monetary saving from pre- 
heating feed-water with .... 213 

piping 363 

portion utilized for feed-water 

heating 229 

separation in vacuum, how 

facilitated 398 

separator, definition 388 

separators 385-401 

supplied by auxiliaries 231 

temperature, feed-water 

heater 224 

Exhaust-head, definition 399 

pollution of atmosphere pre- 
vented by 399 

purpose 399 

Expansion in pipe of given length, 
pipe required for bend, 

formula 378 

in steam separators 389 

linear, coefficient of 378 

in steam pipes, strains due to 377 
in steel and wrought-iron 

steam pipes, formula 378 

stresses in piping systems, taken 

up by bends 370 

F 

Fan pumps 117 

Fan-blower height required for 

cooling tower 361 

Fairbanks-Morse multi-stage cen- 
trifugal pump, illustration 113 
Farnsworth boiler feeder, illustra- 
tion 204 

Feed-pump, boiler, connection to 
feed-water heater, illustra- 
tion 3 

boiler, mechanically-driven, 

illustration 182 

capacity, forcing boilers 197 

constant-speed, water-relief 

valve on 201 

locate close to heater 245 

motors for driving 184 

suction connection 2 

Feed- water, chemical treatment 318 

cold, boiler strain due to 160 

waste of fuel by 207 

cost of treating ...;•_ 319 

formula for determining require- 
ments 195 

fuel saving due to preheating. . . 211 

heater 207-249 

as a purifier 238 

as protective measure 228 

atmospheric 215 

back-pressure valve 220 

Blake-Knowles 233 

centrifugal pump in connec- 
tion with 136 



432 



INDEX 



Page 
Feed-water heaters, classification 

table 215 

cleaned regularly 248 

closed 211 

advantages and disadvan- 
tages. 244 

classification 238 

installation and operation. . 247 

National coil type 238 

piping 247 

safety valve 247 

tubes in 240 

type, table of general 

data 236, 237 

Cochrane, illustration 226 

open induction 219 

connection to boiler feed- 
pump 3 

counter-current, diagram 239 

diagram 239 

double installation 244 

economies 211 

exhaust steam, definition 210 

temperature 224 

function 224 

heat transmission table 243 

heater horsepower rating 243 

heating-surface, formula 242 

horizontal closed 220 

induced or draw heaters 220 

induction, piping of 223 

induction-type open, piping, 

illustration. 222 

installing primary and sec- 
ondary, illustration 216 

intermittent delivery of ex- 
haust steam 242 

open 210 

advantages and disad- 
vantages 244 

explanation 225 

installation 245 

size of shell required 234 

type, table of general data 235 

typical installation 221 

operation 246 

pan or tray area required .... 232 

piping 2 

arrangements 220 

pressure 215 

primary 215 

and secondary, operated 

alternately 218 

application 216 

purity of water delivered to 

by surface condenser 318 

reason used 207 

savers of coal 214 

secondary 215 

open or closed type 217 

selection 238 

steam condensed by, formula. 231 

steam-tube 238 

application 241 

illustration 239 

Stilwell through-type, piping 

arrangement 223 

through or thoroughfare 

heaters 220 

used with injector 160 

vacuum 215 

and atmospheric, in con- 
densing plants 217 

water-tube 238 

heating equipment, classes 210 

saving due to computed 
graphically 213 



Page 

Feed-water, impurities 227 

live-steam heaters and purifiers 248 

oily 228 

per cent, returned to boiler in 

surface condenser 319 

preheating with exhaust steam, 

monetary saving resulting. . 213 
quality, factor in condenser 

selection 318 

quality for economizer 273 

requirements, power plant, basis 

of steam consumption 196 

of power plant, estima- 
tion 194-197 

temperature gain, in economizer, 
ratio to loss of combustion- 
gas temperature 266 

temperature for economizer .... 273 

raised, formula 228 

with surface-condenser used re- 
peatedly 288 

Filtering material, packing 246 

Filtration in open heater 225 

Fire-pump, duplex, illustration 27 

pumps for 135 

underwriters' 59 

Fire-insurance underwriters, pumps 

required by 65 

Fisher pump governor 200 

illustration 198 

Fiske, R. A., on turbine drives 133 

on centrifugal pump advan- 
tages 134 

single impeller pump 120 

squirrel-cage induction motor, 

centrifugal pump drives .... 132 
Fittings, extra heavy cast-steel, 

illustration 365 

malleable iron, illustration 364 
frictional-resistance in, table. . .. 15 

to water-flow through 14 

low-pressure cast-iron, illustra- 
tion 365 

right-angled, in steam pipes, 
pressure-drop due to, for- 
mula 377 

standard cast-iron, illustration 364 
malleable iron, illustration 364 
Flanges, companion, methods of 

securing to pipe ends 369 

cost, according to material 369 

methods of attaching to pipe 

ends 368 

Flexibility of boiler operation due to 

economizer 272 

Float connections in open heater .... 246 
Flow through valves in piping, fric- 
tion due to 6 

Flue-gas temperature for econo- 
mizer 273 

Foot-valve in suction line 142 

leakage, effect of 67 

priming ejector use with and 

without 148 

Force, centrifugal, definition 101 

illustration 102 

Forcing boilers, feed-pump capacity 197 

Formulas, reciprocating pump 34 

Foster, D. E., steam pipe sizes, 

graph 374 

Foundation bolts for centrifugal 

pump 140 

for a centrifugal pump 140 

Fountain, see also Spray foun- 
tains 347-350 

spray, with cooling pond 343 

Friction due to flow 6 



INDEX 



433 



Page 
Friction in pipe, effect on pump-suc- 
tion 1 

in straight pipes 5 

mechanical, in pump mechanism 30 

of liquid entering 6 

of water in straight pipes 9 

Friction-head, definition . 4 

in centrifugal pump 130 

of water in cast-iron pipe, table 

of 12 

in straight pipes, table of . . . . 10 

on a pump 5 

on a pump, definition, for- 
mula 6 

Frictional resistances, what included 25 

Fuel economizers 251-275 

see also Economizers. 
heat of, in gases of combustion, 

table 251 

saved by heating feed-water, 

graph 213 

savine due to economizer 272 

formula . 270 

due to preheating feed- water. 211 
waste of by cold feed-water .... 207 

Fulton governor, illustration 199 

on direct-acting boiler-feed 

pump 178 

pump governor, explanation. . . . 198 

Funnel, non-splash 167 

Funneled inlet-orifice, example 17 



Gage glass, shielded from steam- 
temperature fluctuations, 

illustration 401 

water, on steam separator 401 

Gage-pressure, net, how determined . 28 

Gallons per minute, formula for 34 

Gas in feed-water 228 

Gases, exit-temperature in econo- 
mizer. . 269 

relative flow in economizer 265 

Gear drive preferable to belt drive 185 

Gearing, inspection in economizer. . . 274 

" Gebhardt's Steam Power Plant 

Engineering," advantages 

and disadvantages of open 

and closed feed-water 

heaters 244 

table of engine econ- 
omy 287 

list of heater manufacturers. ... 215 

on heat transfer 243 

" Gillette and Dana," on separator 

costs 401 

Gland, condenser tube 302 

" Goulds Manufacturing Company's 
Catalogue," on boiler-feed- 
pump sizes 193 

on motor ratings, deep-well 

pump 97 

centrifugal pump material 101 

open-well pump, illustration ... 83 

pump capacities 193 

rotary pump 152 

table 2 

table of boiler-feed centrifugal 

pump capacities 194 

thrust bearing 121 

Governor, boiler-feed pump, on 
direct-acting steam 

pumps 198-201 

28 



Page 
Governor for turbine-driven pumps, 

discharge-pressure, control . 133 

pump, horizontal type 199 

on turbine-driven centrifugal 

pumps 201 

troubles of 201 

Gravity apparatus or return traps 

for boiler-feed 171 

Grease on outside of condenser tubes, 

how removed 316 

Greene Economizer Company boiler 

heating-surface charts 268 

table of economizer temperatures 

obtained 264 

showing heat of fuel in gases 

of combustion 251 

temperature charts 269 

Green economizer, illustration 253 

Gridiron separator . . 389 

operating principle 394 

Ground-area required, spray-fountain 

ponds 350 

Guide bearing, centrifugal pump, 

illustration 122 

coupling for deep-well pumps. . . 88 
vanes 109 

H 

Hampton Mills, East Hampton, 
Massachusetts, installation 

of economizer 253 

Hancock inspirator, illustration 157 

Hanger, counter-balancing, for steam 

piping 380 

Hangers, plain, for steam piping .... 380 
" Harding and Willard, Mechanical 
Equipment of Buildings," 
data on feed-water heaters 235 
" Harrison Safety Boiler Works Cata- 
log," condenser economy 

graph 286 

condenser saving 282 

separators 387 

Head, dynamic, or pressure, defini- 
tion 5 

exhaust-, definition 399 

friction, definition 4 

on a pump 5 

inlet static, for boiler feed- 
pumps 4 

maximum, impellers for 119 

measured, definition 5 

in pump operation .......... 7 

necessary to overcome frictional 

resistance 14 

of water, converting to unit 

pressure, formula 4 

pressure, and speed of impeller, 

centrifugal pump 108 

static, of a liquid, illustration. . . 4 
of fluid column, definition. ... 3 
total measured, on pump, defini- 
tion, formula 9 

on pump, definition, for- 
mula 9 

pumped against 106 

velocity, definition 4 

Headers, duplicate main, boiler con- 
nection to 372 

Heat absorbed by water in in- 
jector 162 

abstracted from steam by cooling 
water in surface-condenser, 

formula 302 

balance, automatic exhaust 

steam 181 



434 



INDEX 



Page 

Heat extracted in evaporation 340 

insulating for closed heater 246 

latent, of vaporization, in 

water cooling 329 

loss, excessive, from steam pipes, 

how prevented. 382 

from bare steam pipes, con- 
densation due to, formula 381 
and insulated steam pipe. . . 381 

in non-condensing plant 209 

of fuel in gases of combustion, 

table 251 

pump-duty on basis of, for- 
mula 33 

saved by preheating boiler feed- 
water 212 

transfer coefficient in closed 

heaters 242 

in economizer 269 

transference coefficient, value 
effected by conditions in 

condenser tubes 304 

in surface condensers 300 

condenser, table of coeffi- 
cients 304 

transmission, feed-water heaters, 

table 243 

Heater, closed, non-condensing plant, 

duplex boiler feeder in 204 

exhaust, effect on boiler feeding 

efficiency , 174 

feed-water, see Feed-water 

heater 3 

closed, advantages and dis- 
advantages 244 

open, advantages and disad- 
vantages 244 

steam trap for, illustration . . . 405 

used with injector 160 

formula for raising temperature 

of feed-water 228 

horsepower rating, feed-water 

heaters 243 

live-steam, for feed-water 248 

Heaters, feed-water 207-249 

Heater-tube, corrugated 241 

Heating coils, steam traps for 409 

exhaust-steam, heat balance for 182 
system, gravity, condensate 

from 227 

Heating-surface, additional boiler, 

compared to economizer . . . 267 
boiler, temperature difference 

for 268 

economizer, least temperature- 
difference for 269 

feed-water heater, formula 242 

Hershey Chocolate Company's Plant, 

illustration xii 

Holly steam-loop for draining high 

pressure piping, illustration 382 
Hoppes feed-water heater, illustra- 
tion 221 

live-steam heater 232 

purifier installed, illustration 248 
reverse-current exhaust steam 

separator, illustration 390 

Horsepower delivered by injector 163 
heater, feed-water heater rating 243 

of pump 25 

required for pumping, formula. . 32 
total driving, steam pump, 

definition 29 

water, developed by pump, 

formula 25 

water or hydraulic, indicated, 

definition formula 29 



Page 
Hot water handled with centrifugal 

pumps 137 

pumping 2 

returns, feeding of boilers 

with 202 

Humidity, relative average, table. . . 331 
corresponding to wet- and 
dry-bulb temperature dif- 
ferences, table 335 

definition 332 

effect on water cooling 330 

of atmospheric air, how deter- 
mined 332 

Hydraulic efficiency, data necessary 

to determine 28 

of pump, formula 27 

or water horsepower developed 

by pump, formula 25 

packing for pumps 42 

I 

Impact or baffle-plate separator 389 

Impeller, centrifugal pump, theo- 
retical speed 106 

closed-type, illustration 119 

enclosed 119 

for maximum heads 119 

forces which tend to unbalance 114 

Jaeger method of balancing. . . . 115 

open 117 

r.p.m. and velocity of periphery, 

formula 108 

speed, effect of change in 129 

of, and pressure head, cen- 
trifugal pump 108 

Impurities scale inside of boiler 207 

scale-forming 248 

Indicator card, steam pump 28 

diagram, direct-acting steam- 
pump 25 

Infiltration of air through economizer 

setting 260 

Injectors 155-170 

advantages 159 

application 173 

in absence of heater 173 

approximate equation 161 

as a pump, inefficient 172 

becoming hot 172 

breaks. 170 

capacities and weights, table 166 

different types, applications .... 160 

disadvantages 160 

double-tube 157 

efficiency 174 

elementary, illustration 156 

essential parts 156 

factors influencing performance . 163 

failure to lift water 169 

fed from overhead tank, illustra- 
tion 168 

for boiler-feed 171 

horsepower delivered by 163 

inspirator type, piped for 

boiler feeding, illustration. . 173 

installating 166 

lifting 157 

limitations for boiler-feeding 172 

measure of economy 162 

Metropolitan Model-0 159 

non-lifting 157 

not delivering water to boiler. . . 169 

operating and starting 169 

Penberthy automatic, illustra- 
tion 158 

piping of, illustration 167 



INDEX 



435 



Page 

Injectors, positive, operation 159 

principle of, illustration 155 

selection 165 

self adjusting 158 

single-tube automatic, restart- 
ing 157 

special, for exhaust steam 155 

steam at overflow 170 

steam pressure for 172 

suction-pipe strainer 1G8 

theory 156 

troubles 169 

type for given service 165 

Inlet static heads for boiler feed- 
pumps 4 

Inlet-orifice, funneled, example 17 

Inspection of economizers. 274 

Inspirator, Hancock, illustration .... 157 
type injector, piped for boiler 

feeding, illustration .... 173 

Instruments, economizer fitted with . 259 

Insulation for steam pipes 381 

Intake-pipe to suction- well 9 

International Text Book Company, 

table of engine economy . . . 287 

Iron, cast, for steam piping 363 

malleable, for steam piping 363 

wrought, for steam piping 363 

Irrigation, pumps for 135 

J 

Jacobus, D. C, on economies of 

boiler feeding 174 

Jaeger method of balancing impeller 115 
of impeller balancing, illustra- 
tion 116 

Jet condenser, see also Condenser, 

jet . 289 

elementary, volumes of air, 

water and steam, diagram 283 
high-vacuum, with turbine, 

illustration 305 

Johns electrically-driven pump 188 

Joints, condenser, index to condi- 
tion 316 

corrugated expansion, illustra- 
tion 378 

double-slip expansion, illustra- 
tion 378 

double swing or swivel, taking 
up expansion in pipe lines, 

illustration 377 

expansion slip- 377 

flanged, in steam piping 368 

screwed, in steam piping 368 

steam-piping, types used in 368 

Josse, Professor, tests on heat trans- 
ference coefficient 304 

K 

Kansas City Light and Power Com- 
pany economizer installa- 
tion 254 

Kelley, H. H., on condensers 312 

Kent's "Mechanical Engineers' 
Pocketbook " on economies 

of boiler feeding 174 

on open heater tray area 232 

Kieley expansion steam-traD, illustra- 
tion 404 

pump governor, illustration. . . . 200 

Kinghorn pump valve, illustra- 
tion 46 

Kneass' " Practice and Theory of the 

Injector " 164 



Page 

Koerting multi-jet ejector condenser, 

illustration 293 

multi-spray nozzle 344 

roof space for spray cooling, 

illustration 346 

Kroeshell Brothers Company, 

Chicago, power plant 208 

L 

Lap-weld, strength of 367 

Lap-welded pipes 365 

Law of freely falling bodies 105 

Leakage due to cold water in boiler 207 
of air in condenser, prevention 

of 312 

Leathers, pump-plunger, mold for 

forming 99 

LeBlanc surface condenser 311 

Lift, suction, of centrifugal pump, 

how measured 8 

Liner for piston pump 42 

Liquid, corrosive, centrifugal pump 

for moving 151 

pumping 93 

entering, friction 6 

volatile non-corrosive, pumping. 93 
Live-steam heater for feed-water. . . . 248 

piping 363 

purifier for feed-water 248 

separation, economy 386 

purposes of 386 

separators 385-401 

definition 385 

efficiencies table 396 

Load, character of, for economizer 273 
on pump, practice in determin- 
ing. . 28 

Loew absorption exhaust-steam sepa- 
rator, illustration 395 

Loop header system for steam piping 371 
Holly, returning condensation in 

steam piping to boilers 3S3 

Loss by incorrect valve-stem adjust- 
ment 61 

due to inefficient boiler-feeding 

pumps 178 

heat, excessive, from steam 

pipes, how prevented 382 

from bare and insulated steam 

pipes, table 381 

steam pipes, condensation 

due to, formula 381 

in non-condensing plant 209 

hydraulic, how obtained 28 

in pump tests 27 

of pump, definition 26 

in power plant, chart . 252 

of water, greater in open than in 

closed cooling tower 355 

Lost-motion, valve-stem, function of 

in steam-pumps 58 

Louvres, Burhorn sheet-metal cool- 
ing tower 354 

in cooling tower 352 

Low pressure steam, table of prop- 
erties 301 

Lubrication of engine cylinder when 

wet steam is used 386 

• 

M 

Magnesia heat-insulating for closed 

heater 246 

Main pipe with branches, size, for- 
mula 376 



436 



INDEX 



Page 

Make-up water, cooling effect 342 

definition 356 

Management, pump 66 

Marck expansion steam-trap for 
feed-water heater, illustra- 
tion 405 

Marks' "Mechanical Engineer's 
Handbook " on boiler feed- 
ing equipment 173 

on centrifugal pump character- 
istics. 126 

on steam pipe insulation 381 

Masher centrifugal horizontal 
steam separator, illustra- 
tion 391 

Massachusetts pump, illustration 101 
Mechanical drive for boiler-feed 

pump 182 

efficiency reciprocating pump, 

definition, formula 30 

Measured head, definition 5 

"Mechanical and Electrical Cost 

Data," on separator costs.. 401 
" Mechanical Engineers' Handbook," 

on steam pipe insulation. . . 381 

" Mechanical Refrigeration " . 332 

Mercury, inches of, conversion to 

pounds per square inch .... 29 

vacuum gage 284 

Mesh separator 389 

operating principle 393 

Metropolitan Model-0 injector 159 

Midvale Machine Company, boiler 

feed test 188 

Mine drainage, pumps for 135 

Mitchell-Tappen System, high-tem- 
perature natural draft cool- 
ing tower test data table. . . 359 
low temperature cooling tower 

test data, table 358 

Mixing chamber, Worthington cool- 
ing tower 352 

Moffat feed-water heater 224 

Moisture carried from boiler as spray 

or bulk of water 386 

in steam from boiler 385 

Momentum in steam separators 389 

Motor, adjustable-speed, for feed 

pumps 185 

aligning with vertical pump shaft 122 
constant-speed, on boiler-feed 

pump 185 

direct-current 132 

for driving feed pumps 184 

rating for deep-well pump, 

formula 97 

slip-ring induction, for cen- 
trifugal pumps 132 

squirrel-cage induction, cen- 
trifugal pump drives, R. A. 

Fiske 132 

varying-speed, for feed pumps . . 185 
Motor-driven centrifugal pumps 

speed variation 151 

Motor-generator in automatic heat 

balance 182 

Mover, steam-driven prime, selection 

of condenser for 325 

Muntz metal for condenser tubes 302 



N 



Nason bucket-float steam trap, 

illustration 403 

Nation coil heater 241 

Newcomen's condensation engine 278 



Page 
Non-condensing and condensing 
operation, steam consump- 
tion with 280 

plant, duplex boiler feeder in 

closed heater 204 

energy in exhaust steam 207 

exhaust steam available 229 

location of separator 399 

steam useful in feed-water 

heating 230 

Non-return trap, definition 403 

Nozzles, see also Spray noz- 
zles . .. 344-348 

for spray fountain size and 

number. 348 

impact, Cooling Tower Com- 
pany 343 

in spray fountain, spacing 348 

intermingled spray from, illus- 
tration 345 

pump discharge, pressure at, 

illustration 26 

single spray, illustration 344 

spray, impact, Cooling Tower 

Company , . . . . 345 

Schutte and Koerting cooling 

pond 338 

steam, of injector 156 

O 

Oil in boiler feed- water 211 

removed from feed-water 227 

Oil-drip connection in open heater 246 

Oil-eliminators, definition 388 

Oil-separators, cost 401 

definition 388 

part of heater 224 

Oiling machinery 69 

Operating costs, jet and surface 

condensers, table 326 

Operation, condensing, advantages . . 288 
Overflows of injector 157 

P 

Packing, cutting down 43 

effect on pump suction 1 

of governor valve stem 201 

pump rods and stems 69 

water-piston 42 

Pan area required, open feed- water 

heater 232 

Pans of open heater, removed and 

cleaned 247 

Parallel-flow in economizer 265 

Parsons vacuum augmenter, illus- 
tration 307 

" Peele's Mining Engineers Hand- 
book," table of pump effi- 
ciencies 31 

Penberthy automatic injector, illus- 
tration 158 

Pipe anchorage, illustration 380 

butt-welded, method of form- 
ing 365 

capacities for saturated or super- 
heated steam, graph. . . . . . . 374 

cast-iron, table of friction- 
heads 12 

connections for injectors, table 166 

delivery, friction-head in 14 

obstruction 169 

discharge, of steam-trap, check- 
valves in 413 

sizes for, formula 18 

velocity through, formula. ... 24 
double extra heavy grade 363 



INDEX 



437 



Page 
Pipe, extra heavy cast iron, for steam- 
piping systems .-•.-• ^64 

cast-steel, for steam-piping 

systems 364 

grade 363 

malleable iron, for steam- 
piping systems. 364 

fittings, grades used in steam- 
piping systems 364 

for induction heater 221 

for mixing chamber 352 

for steam-power plant, grades 

of, table 368 

hanger, illustration 380 

in steam-piping systems classi- 
fied according to construc- 
tion 365 

lap-welded, method of form- 

ing 366 

preferable to butt-welded for 

steam piping 367 

length necessary for bend to take 
up expansion in pipe of 

given length, formula 378 

lengths equivalent in resistance 

to fittings, table of 15 

lines, double swing joint taking 

up expansion in 377 

uncovered, steam condensa- 
tion rate in, table 412 

low-pressure cast-iron, for steam- 
piping systems 364 

main, size formula 376 

pump suction, with square 

orifice .- • • 6 

size for reciprocating engine, 

formula 375 

necessary to deliver steam at 

given rate, formula 375 

saturated or superheated 

steam, graph 374 

spiral-riveted steel 366 

standard cast-iron for steam- 
piping systems 364 

grade 363 

malleable iron for steam- 
piping systems 364 

steam, bare and insulated, heat 

losses from, table 381 

condensation due to heat loss, 

formula and table 381 

excessive heat loss, how pre- 
vented 382 

insulation 381 

linear expansion producing 

strains in 377 

pressure drop due to globe 
valves and right-angled fit- 
tings, formula 377 

sizes determined graphically. . 373 

thickness of covering 382 

steel and wrought, sizes 364 

and wrought-iron, steam, 
linear expansion in, for- 
mula 378 

grades 363 

trade meaning 367 

straight, friction of water in. . . . 9 

riveted steel 366 

wrought-iron or steel, table 

of friction heads in 10 

suction, for condenser, strainer 

in 313 

sizes for, formula 18 

table of friction-heads in 10 

vacuum, centrifugal pump suc- 
tion 145 



Page 
Pipe vent, connecting high-pressure 
trap with apparatus 

drained 413 

induction heater 246 

vibration, devices to prevent 

transmission, illustration. . . 379 

welded wrought-iron 367 

steel 367 

wrought-iron, inside and outside 

diameters 363 

trade meaning 367 

Pipe-bend facilitating steam flow in 

systems 369 

minimum length of tangent. ... 371 

radius 371 

pressure-drop produced by 377 

radii of 370 

resistance to steam flow in 

piping system decreased by 369 
standard, for piping systems. . . . 370 

Pipe-ends, flanges attached to 368 

Pipe-fittings, friction 6 

Piping, discharge, centrifugal pump 145 
engine, steam separators in, 

illustration 385 

exhaust-steam 363 

main, average pressure-drop. . 376 

for boiler feeding 171 

friction due to flow through 

valves 6 

high-pressure, Holly steam-loop 

for draining, illustration 382 

live-steam 363 

main steam, unit system, illus- 
tration 373 

of steam trap 412 

pump, turn made with elbow. . . 6 
with long-radius bend 6 
sharp turn, made with plugged 

tee 6 

steam, floor stands for 380 

high pressure, condensation 
returned to boilers with 

Holly loop 383 

loop header or duplicate 

headers for 371 

materials for 363 

of power plants 363-383 

of power plant, two separate 

systems 363 

plain hangers for ;-.••• 380 

screwed and flanged joints 368 

single header system 371 

supporting devices for 380 

types of joints used 368 

unit group for 371 

vibration due to pulsating 

steam-flow 379 

wall-brackets for 380 

suction, centrifugal pump 141 

system, expansion stresses taken 

up by bends 370 

quantity of condensation- 
water to be trapped, for- 
mula 411 

steam, grades of pipe fittings 

used 364 

Piston balancing system, De Laval 

Steam Turbine Company 117 

pump, effective area 20 

requisite diameter for water 

end, formula 22 

volume swept by, illustra- 
tion 19 

speed, crank-action pump, 

table 92 



438 



INDEX 



Page 
Piston, steam, direct-acting steam- 
pump, formula for diameter. 23 
effect of vacuum, illustra- 
tion 277 

Piston-pumps, discharge-heads 42 

illustration 41 

Piston-speed, high, relation to pump- 
slip 21 

Pitting on surfaces of economizer 274 
Plant, non-condensing, heating sys- 
tem, boiler-feeding equip- 
ment for. 182 

Plunger, pump, dimensions of cup- 
washers, table 99 

effective area 20 

inside-packed, illustration. ... 20 
requisite diameter for water 

end, formula 22 

rods, deep-well pump, head- 
pressure equivalents, table 98 
Plunger-pump, belt-driven single 

acting 78 

discharge heads 42 

outside end-packed 41 

pump, illustration 40 

Plunger- valve, deep-well pump 88 

Pond, cooling, condensing water 

cooled by 329 

simple, requisite area com- 
puted 342 

spray-fountain, ground area 

required 350 

Pond-area, requisite total 341 

Pot- valves, illustration 45 

Pot-valve type of pump valve 47 

Power developed by condensing and 

non-condensing engine 279 

horsepower of pump 25 

required for pumping, for- 
mula 32 

total driving, steam pump, 

definition 29 

water or hydraulic, indicated, 

definition formula 29 

"Power," on glass water-gages 401 

on pipe trade meanings 367 

power house drawing from. . . iv 
Power plant auxiliaries and acces- 
sories, illustration 208 

condensing steam, non-con- 
densing steam-driven aux- 
iliaries 180 

" Power Plant Engineering," on boiler 

feeding 205 

on economizer cost 275 

R. A. Fiske, on centrifugal pump 

advantages. 134 

Power plant, estimating feed-water 

requirements 194-197 

feed-water requirements on 
basis of steam consump- 
tion 196 

losses in, chart • 252 

steam, grades of pipe for, 

table 368 

piping 363-383 

Power pump, advantages and dis- 
advantages, table 189 

belt-driven 82 

chain-driven. . . 82 

driven by gearing 82 

duplex 82 

rates of discharge 90-92 

simplex 82 

triplex 82 

required to drive centrifugal 
pump at any speed 130 



Page 
.Power pump, to operate spray foun- 
tain 350 

requirements of jet compared to 

_ surface condenser 322 

saving due to condenser, for- 
mula 281 

to remove air and water from 

condenser 284 

Power-factor-correcting synchro- 
nous-motor-driven centrifu- 
gal pumps 132 

"Practical Engineer," on cost of cool- 
ing tower 361 

"Practical Heat" 329 

Preheater in Badenhausen boiler. . . . 255 
Pressure, absolute, in condenser, 

formula 285 

barometric, effect on condenser 285 
draft, in chimney, in inches of 

water, table 263 

drop, average, in exhaust-steam 

main piping 376 

due to globe valves and right- 
angled fittings in steam 

pipes, formula 377 

in steam mains, allowed in 

practice 376 

prevented by receiver- 
separator 388 

produced by gate valves and 

pipe bends 377 

saturated or superheated 

steam graph 374 

friction, definition 4 

head, definition 5 

head and speed of impeller, cen- 
trifugal pump, formula .... 108 

due to in the vessel 8 

inlet, for boiler feed-pumps 4 

measured, definition 5 

of exhaust, selection of separator 

affected by . . . 400 

producing velocity in a pipe. ... 5 
pump intake, at different tem- 
peratures, graph 2 

unit, converting head of water 

to, formula 4 

velocity, definition 4 

water-vapor, in air, how deter- 
mined 336 

working, increased by condenser 277 
wrought iron and steel pipe, 

data. 367 

Priming centrifugal pumps 146 

methods of 147 

ejector 148 

use with and without foot- 
valve 148 

excessive, prevented by receiver 

separator 388 

Priming-pump for centrifugal pump . 147 

with low suction-lift 149 

Proper pump, selection of 93-97 

Proportions of cooling tower, method 

of computing 360 

Psychrometer, sling, to determine 

relative humidity 332 

Pumps, see also Duplex-pumps and 
simplex-pumps . 
see also Plunger-pumps and 
piston-pumps. 

actual discharge of 21 

work of, definition 24 

air below suction-valves, effect 

of 66 

between suction and discharge 
decks, effect of 67 



INDEX 



439 



Page 
Pump, Alberger hurling-water air, 

illustration . . . 307 

and receiver, combined 202 

and suction pipe, passage of 
water through, as a hydrau- 
lic loss 27 

apparatus for replenishing air- 
chamber, illustration 49 

artesian-well 85 

belt-driven single-acting, illus- 
tration 77 

blowing out steam cylinders. ... 68 

boiler-feed, cost of operation 179 

economical 179 

electric-drive for 182 

mechanical drive for . 182 

mechanically-driven, capacity 

of 186 

constant-speed economy . 185 

motor-driven, illustration. .... 177 

motor- or power-driven, in 

non-condensing plant. ..... 179 

steam-driven reciprocating, in 

every plant 179 

steam-piston 65 

table of capacities . . 193 

Burnham compound simplex, 

illustration 64 

calculations 1-37 

capacity at low speed 66 

causes impairing efficiency 66 

centrifugal, see also Centrifugal 
pumps. 

advantages of 134 

and rotary 101-153 

as condenser auxiliary 307 

boiler-feed, illustration 177 

characteristics, graph. . 125 

efficiency for boiler feeding. . . 192 

methods of priming 147 

suction lift of, how measured . 8 

circulatory, for condenser 305 

lower than discharge level of 

condenser 320 

classes of, used with condenser 305 

cleaning out steam piping 68 

combination high-service and 

low-service belt-driven 77 

compared to steam traps for 

boiler feeding 205 

compound condensing, duty and 

steam consumption, table . . 72 

deep-well . 87 

direct-acting 65 

starting 67 

condensate for condenser 305 

condenser, electric or steam, type 

of drive for 322 

crank-action, see also Crank- 
action pumps 75-99 

power 75 

suction and discharge, graphs 91 

crank-and-fly-wheel 75 

application 95 

operation 76 

deep-well, Chippewa, illustra- 
tion 86 

details of 87-89 

illustration 84 

motor rating for, formula .... 97 
plunger rods, head-pressure 

equivalents, table 98 

direct-acting boiler-feed, with 

Fulton governor 178 

feed, efficiency 174 

simple, duty and steam con- 
sumption, table 71 



Page 
Pumps, direct-acting, steam, see also 

Steam pumps, direct-acting39-73 

as condenser auxiliary 305 

boiler-feed pump gover- 
nors 198-201 

for boiler-feed service, selec- 
tion of 65 

indicator diagram 25 

maximum mechanical effi- 
ciency 30 

requisite steam-piston diam- 
eter, formula 23 

steam-driven, application for 

boiler feeding. 180 

discharge pipe, sizes for, for- 
mula 18 

discharging into reservoir 17 

displacement units of 19 

double-acting, computations. .. . 34 
duplex crank-and-fly-wheel. . . 76 

simplex, application 95 

suction, illustration 1 

triplex, illustration 81 

double-suction, balancing of . . . . 117 

dry-vacuum, illustration 79 

or air, for condenser 305 

Wheeler. 310 

duplex, adjustment of steam- 
valve 57 

compared with simplex- 

pumps 64 

cross-heads 62 

definition 53 

displacement of, example 20 

double-acting power, applica- 
tion 96 

fire, illustration 27 

illustration 54 

requisite length for steam- 
valve rod 57 

steam-valve 55 

electric, power and steam, ad- 
vantages and disadvantages, 

table 189 

electrically-driven, advantages. . 94 
failure to catch water due to 
leakage of valves in suction 

chamber 67 

feed, saving by substituting 
electrically-driven for steam- 
driven 187 

feed-water, table of applications 190 

for boiler-feed 171 

for liquids other than water .... 93 
for water-service in buildings ... 59 

friction in valves 6 

frictional resistance offered by 
internal passages and 

valves 14 

friction-head, definition, for- 
mula 6 

friction-head on 5 

Goulds, open-well, illustra- 
tion 83 

governor, horizontal, illustration 199 
on turbine-driven centrifugal 

pumps 201 

troubles of 201 

governor-controlled duplex, illus- 
tration 59 

high pressure, air-chamber 

charging-apparatus for 51 

horizontal double-acting suction, 

arrangement of valves 47 

hot-well 305 

hurling-water, as condenser aux- 
iliary 307 



440 



INDEX 



Page 
Pumps, hydraulic efficiency of, for- 
mula. ....... 27 

losses, definition 26 

or water horsepower developed 

by, formula 25 

hydro-centrifugal, as condeDser 

auxiliary 307 

incorrect adjustment of valve- 
stem as source of loss 61 

inefficient boiler-feeding losses 

due to 178 

inlets connected to two or three 

sources 171 

inside-packed, illustration 41 

installing 68-69 

jet, as condenser auxiliary 308 

load on, practice in deter- 
mining 28 

management 66 

manufacturer, data furnished 

to . 137 

Massachusetts, illustration ..... 101 
mechanically-driven, for boiler 

feeding 176 

modern applications 97 

motor-driven, for boiler feeding. 176 

net work of, definition, formula. 24 

new, running . 66 

of jet condenser, cost of main- 
taining 323 

of surface condenser, cost of 

maintaining 323 

oiling 69 

operation, measured heads in. . . 7 
outside plunger-pump, illustra- 
tion 40 

piston, high vacuum, how 

secured 310 

illustration 41 

or plunger, discharge of, 

formula 22 

volume swept by, illustra- 
tion 19 

plunger, displacement of, for- 
mula 19 

inside-packed, illustration. ... 20 

or piston, effective area 20 

power, see also Power 

pumps 82-97 

power feed, boiler-feeding 

efficiency 176 

power-driven, total efficiency. . . 31 

power-plant, suction piping 40 

pressure at discharge nozzle, 

illustration 26 

priming, explanation 67 

raising water when empty 67 

reciprocating compound, duty 
and steam consumption, 

table 72 

discharge velocity, formula. . . 24 

displacement of 19 

formulas 34 

indicated efficiency, formula. . 26 

indicator diagram 25 

mechanical efficiency, defini- 
tion, formula 30 

net suction-lift, definition .... 2 

rods and stems, packing of 69 

rotary, action of 152 

advantages and disadvan- 
tages 153 

application 153 

definition, illustration 152 

rotative or crank-action, as con- 
denser auxiliary 305 

run continuously 69 



Page 

Pumps, runners, life of 323 

set below suction supply, illus- 
tration 139 

simplex, compared with duplex 

pumps 64 

definition 53 

illustration 54 

length of stroke, explana- 
tion 57 

single impeller, R. A. Fiske 120 

single-acting duplex, applica- 
tion 95 

triplex, diagram 92 

illustration 81 

snifter for replenishing air- 
chamber, illustration 50 

stand-by, direct-acting 192 

starting 70 

steam, duty of, definition 32 

end warmed up 70 

total efficiencies, table 31 

steam-driven crank-and-fly- 

wheel, illustration 75 

for boiler feeding 176 

total efficiency 31 

stopping 70 

submerged-piston 47 

suction lifts at various altitudes, 

table 2 

suction-pipe, funneled end, illus- 
tration 6 

sizes for, formula 18 

tests, hydraulic' losses 27 

theoretical discharge of 21 

water lift at different tempera- 
tures, graph 2 

total driving horsepower, defini- 
tion 29 

efficiency, definition, for- 
mula 30 

values 31 

head, definition, formula '9 

measured head, definition, for- 
mula 9 

triplex double-acting, applica- 

. tion. . 96 

single-acting power, applica- 
tion 96 

turbine, definition 109 

types used as condenser aux- 
iliaries 305 

vacuum-chamber, function of . . . 51 
Vaile-Kimes single-acting deep- 
well 94 

vertical duplex, boiler feed, illus- 
tration 55 

volumetric efficiency, definition, 

formula 22 

volute, definition 109 

water run through when 

stopped 67 

water-level in air-chamber 50 

wet vacuum, for condenser 305 

Wheeler-Edwards combined 

condensate and air 309 

work, in horsepower 26 

Pump-duty, basis of heat consumed, 

formula 33 

of steam consumption, for- 
mula 33 

Pump-piston, canvas-packed illus- 
tration 43 

canvas packing rings for 43 

metal packed, illustration 42 

water-packed, illustration 42 

Pump-plungers, deep-well 86 

dimensions of cup-washers, table 99 



INDEX 



441 



Page 
Pump-plungers for deep-well, illustra- 
tion 87 

leather cup for packing 88 

or piston, requisite diameter for 

water end, formula 22 

Pump-slip affected by high piston- 
speed 21 

average values 21 

definition 21 

explanation 21 

negative 21 

percentage, formula 21 

relation to volumetric efficiency 22 
Pump-suction, effectiveness in lifting 

water 1 

Pump-valve, ball, illustration 45 

bronze disk type, illustration ... 44 

conical-seated, illustration 44 

effective area of opening 48 

flat-seated, illustration 44 

in power-plant pumps 43 

Kinghorn, illustration 46 

seats of metal-disc 45 

securing into valve deck 46 

stems and piston rods, causes of 

scoring .•■••: 70 

used for clear liquids, illustra- 
tion 46 

Pump-work, useful, illustration 7 

Pumping engines 79 

head, jet-condenser circulating 

pump 321 

surface-condenser circulating 

pump 321 

horsepower required, formula. . . 32 

of hot water 2 

unit, selection of 138 

Purification of feed-water by open 

heater 224 

Purifier, feed- water heater as 238 

live-steam, for feed-water 248 

R 

Radiation, cooling by 329 

Radojet, Wheeler two-stage air 

pump, illustration 308 

Rain type, Wheeler low-level jet con- 
denser 294 

Rayne's formula for computing pipe 

length required for bend . . . 378 

Receiver and pump, combined 202 

Receiver-separator, Cochrane hori- 
zontal, illustration 387 

definition 387 

Reciprocating engine condenser prac- 
tice 285 

pipe size for, formula 375 

pumps, compared to centrifugal 

pumps 134 

compound condensing, duty 
and steam consumption, 

table 72 

duty and steam consump- 
tion, table 72 

displacement of 19 

formulas 34 

indicated efficiency, formula. . 26 
mechanical efficiency, defini- 
tion, formula 30 

net suction-lift, definition .... 2 

principle of, illustration 40 

simple, duty and steam con- 
sumption, table 71 

Recooling, atmospheric, of conden- 
sing water 329 

condensing water, methods. 329-361 



Page 
Recooling, effected in cooling tower, 
per cent resulting from evap- 
oration 352 

in spray fountains conditions 

affecting 344 

system, devices for bringing air 

and water into contact 337 

Relative humidity, definition 332 

effect on recooling 345 

Resistance due to water friction 9 

to steam flow in piping system 

decreased by pipe bends . . . 369 

Return trap 202 

boiler-feeding 203 

definition 403 

or gravity apparatus for boiler- 
feed 171 

Reverse-current separator 389 

operating principle 389 

Rings, canvas packing for pump pis- 
ton 43 

metallic, packing for pumps. ... 42 
snap, water-piston packed with, 

illustration 39 

pipes 365 

steel pipe, spiral 366 

straight 366 

Rods and stems of pump, packing 

of 69 

steam-valve, duplex-pumps, 

illustration 58 

piston, causes of scoring 70 

Roof, power-house, spray-fountains 

on 350 

Roof-space for spray cooling, illustra- 
tion 346 

Rotary pump, definition, illustra- 
tion 152 

Royal Technical School, Charlotten- 
burg, tests on heat transfer- 
ence coefficient 304 

S 

Safety valve, closed feed-water heater 247 

inspection in economizer 274 

Saving by substituting electrically- 
driven for steam-driven feed 

pump 187 

due to feed-water heating com- 
puted graphically 213 

monetary, resulting from pre- 
heating feed-water with ex- 
haust steam 213 

power due to condenser, for- 
mula. 281 

Scale forming in economizers . 259 

inside boiler from impurities. . . . 207 

Scale-forming impurities 248 

Schutte and Koerting Company, 

spray-nozzle capacities table 348 
table of spray-fountain data. . . . 349 
double spraying system, dia- 
gram 338 

straight-tube closed heater 240 

cooling guarantees 347 

Scraper, economizer-tube 257 

tube, power expended 258 

Screens for mixing chamber ........ 352 

Sealing surfaces, fit between 119 

Seats of metal-disc pump-valves .... 45 

Sediment in economizer-tubes 259 

Sellers' injector, self-adjusting num- 
ber 8, performance, graph 164 

Restarting Injector 165 

Separating-plate, Bundy steam sepa- 
rator, illustration 394 



442 



INDEX 



V 



Page 
Separation efficiency and velocity of 

steam flow graph 397 

exhaust-steam, economy 388 

in vacuum, how facilitated . . . 398 

purposes 388 

live-steam, economy 386 

purposes of 385 

maximum efficiency attainable. . 395 

Separator, absorption 389 

and appurtenances for efficiency 

test, illustration 397 

Austin reverse-current live- 
steam, illustration 390 

centrifugal 389 

oi^rating principle 390 

efficiency affected by velocity of 

steam-current 396 

exhaust-steam, definition. 388 

Loew absorption 395 

proper location for 399 

selection of 400 

gridiron 389 

Hopptes reverse-current exhaust- 
steam, illustration 390 

- Harrison Safety Boiler Works 387 

impact or baffle-plate 389 

live- and exhaust-steam. . . . 385-401 
in engine piping, illustration . . 385 
live-steam, Austin baffle-plate 

angle, illustration . . 391 

definition 385 

drained automatically 400 ; 

efficiencies table 396 

efficiency of, formula 397 

on basis of steam quality, 

formula... 398 

high^pressure steam trap for, 

illustration 405 

operation and structure 387 

proper location for 399 

selection of 400 

. storage capacity .• • • • 387 

Stratton centrifugal horizon- 

<; tal, illustration-.^. 391 

*witn*jarge well. . .?*' 387 

locations selecti6n affected by . . . 400 

mesh... It : . ..." 389 

oil, cost. .'. ..+...: . ":. 401 

for feed-water 227 

of feed-water heater, illustra- 
tion. . .'. . y. 225 

receiver-, definition 387 

excessive primin prevented by 388 
pressure drop prevented by . . . 388 

steam-storage capacity 388 

vibration prevented by 388 

reverse-current . 389 

operating principle 389 

steam, Bundy gridiron, illustra- 
tion. . 394 

classification 389, 

cost j 401 

functions of corjrugated sur- 
faces i 393 

Masher horizontal centrifu- 
gal, illustration . . 391 

physical phenomena involved 389 

size ../.... 400 

Swartwout/centrifugal, illus- 
tration^. 391 

Sweet ■Vertical, illustra- 
tion .f 393 

with gliFss water-gages 401 

vacuum trap for draining 404 

Weperon reverse-current re- 

/ ceiver-, illustration 390 

well, function 387 



Page 
Setting, economizer, leakage of air 

into 260 

inspection in economizer 274 

Sewage, centrifugal pumps for 

moving 151 

pumping plants, pumps for 135 

Shell, size required for open feed- 
water heater. . .• 234 

Side-suction pump. 114 

Side-thrust in centrifugal pump 114 

Simplex double-acting- pumps, appli- 
cation 95 

pump, air-chamber for 50 

as vacuum- and air-pumps ... 65 
compared with duplex-pumps 64 
single-acting, rates of suction 

and discharge, graph 90 

steam- valve 55 

Single header system for steam 

piping. . 371 

Single-suction pump 114 

Siphon jet condenser . . 289 

Sling psychrometer to determine 

relative humidity 332 

Slip, pump, definition, explanation. . 21 
relation to high piston- 
speed 21 

Smallwood, Julian, "Mechanical 

Laboratory Methods " 161 

Snifter for air-chamber of pump, illus- 
tration 50 

Speed variation, motor-driven cen- 
trifugal pumps 151 

Spray cooling and condenser-outfits, 

performance guarantees. . . . 347 
constant, proper value, pre- 
determination 347 

effect on condensing water . . . 329 
roof space for, illustration. . . . 346 
fountain, conditions affecting re- 
cooling 344 

in connection with cooling 

ponds 343 

installations, table of related 

data 349 

nozzles, size and number 348 

spacing 348 

on power house roof 350 

ponds, ground-area required. . 350 
power required to operate. . . . 350 

nozzle, Badger 344 

capacities, table 348 

impact, Cooling Tower Com- 
pany 345 

installation, temperature re- 
duction effected by, formula 347 

tests, graph 346 

pond for artificial cooling 318 

with Cooling Tower Com- 
pany's impact nozzles... 343 
Springs, pump discharge valves, pres- 
sure to overcome reaction, 

as a hydraulic loss 27 

Stand-by pump, direct-acting 192 

Standoipe, column of water in a. . . . 3 

illustration 4 

Stands, floor, for steam piping. ..... 380 

Static head of fluid column, definition 3 
Steam and oil separators, Direct 
Separator Company, on 

separator economy 387 

at injector overflow 170 

automatic exhaust, heat balance 181 
boilers, apparatus for feeding 

water 171 

condensation rate in uncovered 

pipe lines, table. 412 



INDEX 



443 



207 
209 



Page 
Steam, condensed by open feed- water 

heater, formula 231 

condenser, see also Condenser 

steam 277-327 

condensing water for 329 

consumption, duty of pump on 

basis of, formula 33 

power plant feed-water re- 
quirements based on 19G 

relation to condenser vacuum 286 
with condensing and non-con- 
densing operation 280 

crank-and-fly-wheel pumps 79 

end, pump, warmed up 70 

energy conserved by separation. 386 
exhaust, see also Exhaust steam. 
energy in, non condensing 

plant 

heat in 

main piping, average pressure- 
drop 376 

separating for heating system 388 

separators 385-401 

from boiler, moisture in 385 

heat abstracted from by cooling 
water in surface condenser, 

formula 302 

heating, low-pressure 180 

how saved by condenser 279 

in condensing plant, useful for 

feed-water heating 230 

line tapping for injector 166 

live, separators 385-401 

low pressure, table of prop- 
erties 301 

main, floor stand supporting. . . . 380 
pressure-drop allowed in prac- 
tice 376 

suspending and counterbal- 
ancing expansion loops 380 

wall-bracket supporting 380 

net thermal value diminished by 

moisture 386 

nozzle of injector 156 

pipe size necessary to deliver at 

given rate, formula 375 

Steam piping, see also Piping, 

steam 363-383 

of power plants 363-383 

power plant, grades of pipe for, 

table 368- 

pumps, advantages and disad- 
vantages, table 189 

allowable velocity in water- 
piping 40 

direct-acting 39-73 

classification 41, 53 

for boiler-feed service, selec- 
tion of 65 

hydraulic pressure, illus- 
tration 

outside center-packed, illustra- 
tion 

reciprocating double-acting. . . 
vacuum-chamber connected 

to, illustration 

valve-stem lost-motion, func- 
tion of 58 

with outside end-packed 
plungers, water end, illustra- 
tion 20 

quality, efficiency of live-steam 

separator based on 398 

saturated or superheated, pres- 
sure-drop, pipe sizes and 

capacities for, graph 374 

saving due to condenser, graph . 282 



45 



52 



X^Page 
Steam separators, see also Separators, 

steam 385-401 

cost 401 

supplied by boiler plant, methods 

of distributing 371 

useful in feed-water heating, 

non-condensing plant 230 

volume required for discharge of 

return trap 406 

water in, turbines damaged by 386 



wet, effect on engine cylinder 

lubrication 3: 

loss of turbine efficiency due 

to 

reasons for separating ...*.... 

Steam-bound, pumps becoming .... ..' 

Steam-current velocity through 

afator, efficiency affecte< 

by. .' 

Steam-flow in piping system, 
ance decreased . by 
bends. . 
pulsating, causing y 

Steam piping 
velocity and 
graph . . 

velocities in prac^feeaj ■■?. 375 

Steam-gage pressure J'^ta. balance 
water-gage /pressure, for- 
mula . . . . 4 . . . «;■£!. 36 

Steam-loop, Holly, |for draining- high- - 

pressure piping, illustration 382 
Steam-piston diameter, requisite, 
direct-acting steam'vpump, 

receiver 




**•' 



!p£ :jT...S. 397 . 



formula. 
Steam-storage capacity, 

separator . ...•. . j| . *". . . 

Steam-temperature ,.; fluctuations, 



23 



388 



401 




shielding glass water-gages 
from. 
Steam traps, 
steam. 
"Bulletin" <fc Elliott -Cor 
condensatffn rates, tabjj 
"CatechismjS wendemar 
pacities andl dii 

table ft .',. 

compared to puWps 
feeding ...... !r^ 

separator drainedj 
Steam-valve, adjustment of, duplex- 
pump A .*.' 57 

of duplex-pump/! 55 

rod, dupjex-pump, requisite 

length. ...;..... 57 

Steel, cast, for steam piping 363 

mild, for steam-piping 363 

pipe and wroilght iron, working 

pressures, {data 367 

trade meaning 367 

Stems and rods off* pumps, packing 

of AJ 69 

steam-valve* duplex-pumps, 

illustratiofi 58 

pump-valve, causes of scoring 70 

Stilwell feed-water heater 217 

through-type feed-water heater, 

piping arrangement 223 

Soot, methods of removing from 

economizer-tubes. 257 

Soot-blower, economizer, .z . 258 

Soot-blowing system, steam-con- 
sumption of .* 258 

Soot-pit, inspection in economizer. . . 274 

Soot-scraper, illustration 258 

inspection in economizer 274 

Storage of feed-water in open heater 225 



\ 



444 



INDEX 



Page 
Strainer, foot- valve with, illustra- 
tion 142 

in circulating water suction 

pipe 313 

in trap-inlet connections, func- 
tion 412 

injector suction-pipe 168 

pump suction-pipe, illustra- 
tion 6 

Strains, avoided by preheating 173 

Stratton centrifugal horizontal live- 
steam separator, illustra- 
tion 391 

Stresses, boiler, diminished by use of 

economizer 272 

Striking-points equally spaced, illus- 
tration 62 

Stroke, length in inches, - formula 

for 35 

Strokes per minute, formula for 36 

Stuffing-box, condenser, index to 

condition 163 

Stuffing-boxes of centrifugal pumps, 

packing 151 

Sturtevant, B. F. Company, econ- 
omizer draft fans 262 

functions, diagram 251 

economizer, header and tube 

construction 256 

Suction-lift, net, of reciprocating 

pump, definition 2 

of centrifugal pump 139 

how measured 8 

of pump at various altitudes, 

table 2 

of the water 7 

practical maximum 1 

pump with low, priming 149 

Suction line, enlarged, connected to 

centrifugal pump 143 

imperfectly laid, illustration. . 8 
independent, centrifugal 

pumps 142 

nozzle, pump, pressure in, illus- 
tration 40 

pipe of pump, sizes for, formula 18 
pump, funneled end, illustra- 
tion 6 

rate of, crank-action pump, 

graph 91 

Suction-supply, distant 9 

Suction well for pump supply 9 

illustration 8 

Sump for single centrifugal pump . . . 143 
Surface condenser, see also Condenser, 

surface 289 

coefficients of heat transfer- 
ence, table 304 

feed- water used repeatedly . . . 288 
heat abstracted from steam by 

cooling water formula 302 

tubes and tube-sheets 302 

Surface condensing plant location of 

separator 399 

corrugated, in steam separator, 

functions 393 

external, and internal, inspection 

in economizer 274 

water-cooling, required in surface 

condenser, formula 303 

Swartwout centrifugal steam sepa- 
rator, illustration 391 

Sweating economizer-tubes 257 

Sweet mesh, exhaust-head, illustra- 
tion 399 

vertical steam separator, illus- 
tration 393 



Page * , 
Swendeman's, "A Steam-Trap Cate- 
chism," capacities and di- 
mensions table 411 

Symbols, list xii 

Synchronous-motor-driven centrifu- 
gal pumps, power-factor- 
correcting 132 



Tee, plugged, for sharp turn in pump ^f*' 

piping -.-£ 

Temperature, breaking, of injector. .164 
"drop" in surface condensers, 

definition 304 

effect on pump intake pressures, 

graph 2 

of atmospheric air, how deter- 
mined 332 

of water, effect on pump suction 1 
reduction effected by spray- 
nozzle installation formula 347 
average, effected by cooling 

tower in summer 356 

wet and dry-bulb, table 331 

Temperature-difference for boiler 

heating-surface 268 

least, for economizer heating- 
surface 269 

wet- and dry-bulb, relative 
humidities corresponding to, 

table 335 

Terminal difference, definition 320 

jet condenser 320 

Terry Steam Turbine Company, 

pump test 136 

Testing centrifugal pump, illustration 

and formulas 123 

Tests, pump, hydraulic losses 27 

Thermometer, dry-bulb 332 

economizer fitted with 259 

wet-bulb 332 

Thrust-bearing for vertical sub- 
merged centrifugal pump 152 
Goulds Manufacturing Com- 
pany 121 

Tile-tubing for mixing chamber 352 

Tilting or dumping steam trap, 

counterweighted 406 

Total efficiency of pump, definition, 

formula 30 

Tower, cooling, see also Cooling 

tower 339-361 

Burhorn metallic 330 

condensing water cooled by 329 
Trap, ball-float, valve-operating 

mechanism, illustratiorl .... 407 

Bundy, illustration .,\ 203 

continuous-discharge a- • • • 405 

discharge pipes, check-v&lves 

.in ..!.... 413 

draining discharge pipes . \ •.}. . . . 412 

expansion, location .j. . . . 410 

inlet connections, strainer^ in, 

function . . . .; . . . . 412 

intermittent-discharge ..;.*'.... 405 

return .". I 202 

coal saving effected by. J 408 

operating principle 3 405 

volume of steam required for 

discharge j* ./. 406 

steam :. ,'. . 403-413 

capacity ... ~.%.i 410 

care of : . » v . 413 

classification according to dis- 
charge. . . . i Mj X 405 

to operation . , iV 403 



*. I 








^^ Page 
Trap, steam, compared to pumps for 

boiler feeding 205 

definition .■ : • • ^03 

dimensions and capacities, 

table 411 

expansion, for feed-water 

heater 405 

external by-pass 412 

for heating coils 409 

for live steam separators, illus- 
tration 405 

high- or low-pressure, location 

for 409 

Kieley expansion, illustration. 404 
methods of detecting leaks. . . 413 

non-return, economy 408 

piping 412 

return and non-return 403 

economy 407 

proper location for 408 

temperature-operated, limita- 
tions 409 

vacuum, for draining separator 404 
vent-pipe connecting with 

apparatus drained 413 

Trapping-sheet in steam separator 393 
Tray area required, open feed-water 

heater 232 

Trays, perforated, for mixing 

chamber 352 

Triplex double-acting pumps, appli- 
cation , 96 

single-acting power pumps, 

application 96 

Tubes and tube-sheets, surface 

condenser 302 

condenser, fouling of, result. . . . 315 
economizer, scale and sediment 

in 259 

in closed feed-water heater 240 

surface-condenser, replacement 325 
Tube-cleaner, Worthington hydraulic 324 
Tube-cleaning equipment, built-in, 

surface condenser 324 

Tube-sheets, surface condenser 302 

Tube-surface required in surface con- 
denser, formula 303 

Tube-surfaces, economizer, cleanli- 
ness 257 

Tuck piston-packing, illustration. ... 43 

Turbine condenser practice 286 

vacuum for '. 285 

damage due to water in steam . . 386 
for centrifugal pump, water 

rate 193 

pumps, definition 109 

steam, as pump drive, condens- 
ing 133 

driving centrifugal pump 133 

with high-vacuum jet con- 
denser 305 

Turbo-generator, Westinghouse- 
LeBlanc surface condenser 
with 311 



U 



Uniflow-engine condenser practice. . . 286 

Underwriters, fire-insurance, pumps 

required by 65 

Underwriters' fire-pump, illustra- 
tion 59 

Unit system main steam piping, 

illustration 373 

Unit-group arrangement for steam 

piping 372 



445 



Page 
V 

Vacuum augmenter, Parsons, illus- 
tration 307 

condenser, loss while running. . . 313 
relation to steam consump- 
tion 286 

conversion inches of mercury to 

pounds per square inch .... 29 
corresponding to condenser tem- 
perature 315 

exhaust-steam separation in, how 

facilitated 398 

gage, illustration 284 

high, with piston pumps, how 

secured 310 

in jet condenser, how restored. . 314 
most profitable, in condenser. . . 285 
partial, in condenser, increased 

engine economy 277 

Vacuum-chamber connected to 

steam-pump, illustration 52 
height of water in, illustration . . 52 

in pumps, function of 51 

special, illustration 52 

Vacuum-pump, simplex pump as. . . . 65 
Vaile-Kimes single-acting deep-well 

pump 94 

Valve arranged above pump-barrel, 

illustration 47 

back-pressure, feed-water heater 220 
decks, securing pump valve seats 

in. . 46 

direct-acting simplex steam- 
pump, illustration 56 

discharge, closed, centrifugal 

pump run with 149 

effect of tightness on pump 

suction . 1 

exhaust-relief, back-pressure, in 

water piping 246 

flat disc, area of opening 49 

flat-faced bronze poppet 47 

wing poppet, illustration 46 

flat-seated pump, illustration ... 44 
foot, with strainer, illustration.. 142 
frictional resistance in pump due 

to 14 

table 15 

to water-flow through 14 

gate or check, in pump discharge 

line 67 

pressure-drop produced by . . . 377 
globe, in steam pipes, pressure- 
drop due to, formula 377 

in piping, friction due to flow 

through 6 

inside-operated, simplex steam- 
pump, illustration 56 

leakage, failure to catch water. . 67 
leaky, in steam boiler-feed 

pumps 192 

of horizontal double-acting 
suction-pumps, arrange- 
ment 47 

of power-plant pumps 43 

pump, see also Pump-valves . . . 43-47 
discharge, pressure to over- 
come reaction of springs, as 

a hydraulic loss 27 

friction in 6 

rubber, for low pressure 46 

seat of governor 202 

uneven wear in 68 

side by side above pump-barrel, 

illustration 48 

steam, of simplex-pumps 55 



7- 



446 



INDEX 



Page 
Valve, steam-pump, wear causing lost 

motion ... , 68 

steam-thrown 55 

suction, air below, effect on 

pump... 66 

uneven wear in 68 

water-relief, on constant-speed 

feed pump 201 

Valve-discs, rubber composition, 

hardness 45 

Valve-operating mechanism of Ameri- 
can ball-float steam trap, 

illustration 407 

Valve-orifice of steam trap, area .... 408 

Valve-stem lost-motion, correct 62 

in steam-pumps, function of . . 58 
slow-running duplex-pumps 63 

of governor, packing 201 

pump, incorrect adjustment as 

source of loss 61 

rigid, connection in duplex- 
pump 63 

Vanes, diffusion 109 

guide 109 

Vapor pressure of saturated water 

vapor, graph 336 

water, in air, weight, how deter- 
mined . 336 

pressure in air, how deter- 
mined. . . 336 

Velocity, allowable in water-piping of 

direct-acting steam pump . . 40 

steam-flow, in practice 375 

head, definition 4 

why neglected 5 

of freely falling body, formulas 105 
of periphery and r.p.m. of 

impeller, formula 108 

current through separator, effi- 
ciency effected by 396 

flow and separation efficiency 

graph 397 

in separator selection 400 

of water in cast-iron pipe, table 

of 12 

in pipes, table of. 10 

pressure producing in a pipe .... 5 
through reciprocating pump dis- 
charge pipe, formula 24 

Vent-pipe connecting high-pressure 
trap with apparatus 

drained 413 

induction heater 246 

Vertical pump, aligning motor with 122 

bearings in 121 

Vessel, head due to pressure in 8 

Vibration in steam piping due to 

pulsating steam-flow 379 

pipe, devices to prevent trans- 
mission, illustration 379 

prevented by receiver-separator 388 
Voltage, steady, electrically-driven 

centrifugal pumps 151 

Volume of water and friction head in 

centrifugal pump 130 

swept by pump piston, illustra- 
tion 19 

Volumetric efficiency of pump, defini- 
tion, formula 22 

relation to pump-slip 22 

Volute pumps, definition 109 

Vulcan soot blower 258 

W 

Wainwright closed feed-water heater, 

copper corrugated tube • • < , 240 



Page 
Washers, cup, for deep-well pump- 
plungers 98 

Waste-valve of injector 157 

Water and air, power to remove from 

condenser 284 

apparatus for feeding steam 

boilers 171 

area of in cylinder, formula 

for...... 35 

bad, effect in jet condenser 319 

on surface condenser 319 

column of, converting to unit 

pressure, formula 4 

condensing, see also Condensing 
water. 

recooling methods 329-361 

cooling, character, quantity and 
source factors in condenser 

selection 318 

cost of handling. 320 

devices for bringing into contact 

with air in recooling system 337 
discharged from jet condenser, 

temperature 300 

heads, high, effect on condenser 

water requirement 322 

height raised by pump suction. . 1 

hot, pumping 2 

"make-up," cooling effect 342 

definition 356 

passage through suction pipe 
and pump, as a hydraulic 

loss 27 

pounds pumped per pound of 

steam 162 

pumped per pound of steam 
decreases with steam pres- 
sure 164 

quantity of, delivered by cen- 
trifugal pump 107 

raised to height. 112 

ratio to steam, jet and surface 

condensers 321 

relative flow in economizer 265 

slugs, dangers 386 

vapor in air, weight, how deter- 
mined 336 

pressure in air, how deter- 
mined . 336 

saturated, pressure graph .... 336 
weight of, in one cubic foot of 

air, graph 336 

Water-cooling in surface condenser. . 290 
surface required in surface 

condenser, formula 303 

Water-end, crank-action pumps 80 

of direct-acting steam pump, 

illustration 39 

Water-gage, glass, on steam separator 401 
pressure to balance steam-gage 

pressure, formula 36 

Water-horse-power developed by 

pump, formula 25 

indicated, definition formula. . 29 
Water-loss, cooling tower, per cent. . 357 
greater in open than in closed 

cooling tower 355 

Water-piston, packed with snap rings, 

illustration 39 

packing 42 

Water-relief valve on constant-speed 

feed pump 201 

Wearing rings 115 

illustration 116 

Welderon reverse-current receiver- 
separator, illustration 390 

Well, deep-well pump for 84 



?% 



INDEX 



447 



Page 
Well, driven, centrifugal pump. . .i . j 144 
or sump for single centrifugal \ 

pump 143 

suction, illustration 8 

Westinghouse-LeBlanc surface con- 
denser 311 

turbine with, illustration. . . 306 

Wheeler cooling tower 351 

dry-vacuum pump 310 

rain type low-level jet con- 
denser 294 

two-stage "Radojet" air-pump, 

illustration 30S 

Wheeler-Balcke natural draft 

cooling tower 353 

Wheeler-Edwards condensate and air 

pump 309 

Whitlock closed heater, manifold 241 

Wind velocities, table 331 

Wind-break around spray fountain. . 350 

Wing- valves in high-pressure pumps 45 

Work, actual, of pump, definition . . . 24 

gained by condensing operation, 

diagram 280 



Page 
Work, net, of pump, definition, for- 
mula 24 

of pump in horse-power 25 

Working pressures, wrought iron and 

steel pipe, data 367 

Worthington Company, forced draft 
cooling tower with surface 

condenser 354 

cooling tower, illustration 339 

mixing chamber 352 

forced draft, cooling tower 353 

hydraulic tube-cleaner 324 

independent jet condenser, illus- 
tration 290 

Worthington Pump and Machinery 
Corporation, centrifugal 

pump 110 

thrust bearing 121 

standard condenser-tube gland 302 
Wright baffle-plate exhaust-head, 

illustration 399 

Wrought iron and steel pipe, working 

pressures, data 367 

pipe, trade meaning 367 



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